Hydraulic controller for variable capacity hydraulic transmission

ABSTRACT

The present invention provides a hydraulic controller for a variable capacity hydraulic transmission. This hydraulic controller reduces the load of the engine to improve fuel efficiency while maintaining a hydraulic pressure which is necessary for the control of the skew angles of the swash plates of the variable capacity type hydraulic pump and hydraulic motor that are incorporated in the transmission. The hydraulic controller adjusts the hydraulic pressure supplied to the servo cylinders  92  and  96  which are to tilt the pump swash plate  32  and the motor swash plate  38 , to a minimum necessary pressure on the basis of the pressure Ph of the oil passage which is at the higher pressure in the hydraulic closed circuit  26 . The hydraulic controller adjusts the hydraulic pressure also on the basis of the types of the valve plates of the hydraulic pump  24  and the hydraulic motor  25 , the skew angles of the swash plates, the rotational speed Np of the hydraulic pump and the rotational speed Nm of the hydraulic motor, and the operational condition of the transmission.

FIELD OF THE INVENTION

The present invention relates to a hydraulic controller used for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor. This pump is driven by an engine to deliver hydraulic oil that actuates the motor, which then generates a rotary driving power .

BACKGROUND OF THE INVENTION

Various types of hydraulic transmissions have been proposed each comprising a hydraulic pump and a hydraulic motor, at least one of which is in variable capacity. Pumps and motors of swash plate and plunger type are well known for use as hydraulic variable delivery pumps and variable displacement motors in such transmissions. In addition, there is another type of hydraulic transmission which is known as hydromechanical transmission (e.g., Japanese Laid- Open Patent Publication No. S62(1987)- 147148). This type of transmission is designed to increase the efficiency of power transmission and includes a mechanism for carrying out mechanical power transmission, which is provided parallel to a power transmission mechanism that comprises a hydraulic pump and a hydraulic motor. The above mentioned former type, which comprises only a hydraulic pump and a hydraulic motor, is generally referred to as “hydrostatic transmission”. However, here, in this document, both types, including the latter, are referred generally as “hydraulic transmission.”

There is a well known variable speed control which is used, for example, to control a vehicle that comprises such a hydraulic transmission. In this variable speed control, a value which corresponds to the accelerator opening (i.e., the amount of accelerator operation, engine throttle opening, etc.) is defined as “accelerator index”, and a value which corresponds to the vehicle speed or to the rotational speed of the motor that is proportional to the vehicle speed is defined as “vehicle speed index.” In relation to these values, a target engine rotational speed is predetermined, and the capacities of the hydraulic pump and the hydraulic motor are controlled to bring the actual engine rotational speed to this target engine rotational speed. To control the capacities of the pump and the motor, the respective swash plates of the pump and the motor are tilted, i.e., the skew angles of the swash plates are controlled.

The control of the skew angles of the swash plates of the pump and the motor is carried out by the expansion and contraction of the servo cylinders which are provided to the pump and the motor, respectively. This expansion and contraction of the servo cylinders is effected by appropriating the line pressure of the hydraulic circuit of the transmission. These pressures which are required to actuate the respective servo cylinders vary in correspondence with the moment forces that act on the swash plates of the pump and the motor, respectively, and they become maximum when these moment forces become maximum. Therefore, for the servo cylinders to be effective all the time, the line pressure must be large enough to meet the maximum moment forces.

However, in practice, the moment forces acting on the swash plates rarely reach the expected maximum values. This means that maintaining the line pressure to such high level all the time so as to meet the requirements of the maximum moment forces, which rarely occur, or keeping the line pressure always above a general or normal requirement is wasteful and not fuel efficient because this condition burdens the engine.

SUMMARY OF THE INVENTION

The present invention is to solve the above problem. Therefore, it is an object of the present invention to provide a hydraulic controller which is able to maintain a hydraulic pressure necessary for the control of the skew angles of the swash plates of the pump and the motor that are incorporated in a variable capacity hydraulic transmission.

In order to achieve this objective, the present invention provides a hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, the hydraulic pump driven by a prime motor (e.g., the engine E in the following embodiment) delivering oil that actuates the hydraulic motor into rotation. This hydraulic controller comprises hydraulically actuated swash plate tilting means (e.g., the servo cylinders 92 and 96 in the embodiment) which tilts the swash plate of at least either the hydraulic pump or the hydraulic motor (e.g., the pump swash plate 32 and the motor swash plate 38 in the embodiment) around an axis to control the capacity thereof. The hydraulic controller further comprises pressure controlling means (e.g., the regulator valve 60 and the first linear solenoid valve 51 in the embodiment) which adjusts a hydraulic pressure (e.g., the line pressure PL in the embodiment) used for actuating the swash plate tilting means. The hydraulic controller also comprises estimating means (e.g., the control unit ECU and the various sensors in the embodiment) which estimates a moment force that acts on the swash plate from the above mentioned at least one of the hydraulic pump and the hydraulic motor. In this hydraulic controller, the pressure controlling means adjusts the hydraulic pressure in correspondence with the moment force estimated by the estimating means.

In this way, the hydraulic pressure required for the actuation of the swash plate tilting means is prevented from becoming unnecessarily high. Therefore, the load of the prime motor is reduced, thereby improving the fuel efficiency. Hire, the term “moment force” refers to a periodic force which acts around the trunnion of the respective swash plate while the hydraulic pump or the hydraulic motor rotates with the pistons thereof reciprocating inside the cylinders because of the inside pressures which vary accordingly.

Preferably, the hydraulic controller includes a hydraulic closed circuit which comprises a first oil passage connecting a port of the hydraulic pump (e.g., the port 24 b of the hydraulic pump 24 in the embodiment) with a port of the hydraulic motor (e.g., the port 25 a of the hydraulic motor 25) and a second oil passage connecting another port of the hydraulic pump (e.g., the port 24 a of the hydraulic pump 24 in the embodiment) with another port of the hydraulic motor (e.g., the port 25 b of the hydraulic motor 25). In this hydraulic closed circuit, the estimating means comprises first pressure detecting means (e.g., the pressure sensor 208 in the embodiment) which detects the pressure of the first oil passage and second pressure detecting means (e.g., the pressure sensor 209 in the embodiment) which detects the pressure of the second oil passage, and the estimating means estimates the moment force from the pressure which is higher of the two pressures of the first and second oil passages detected by the first and second pressure detecting means. In this way, the adjustment of the hydraulic pressure for the swash plate tilting means is carried out accurately because the varying moment force affects mostly the pressure of the oil passage which is at the higher pressure in the hydraulic closed circuit, and the moment force is estimated from this pressure.

Preferably, the estimating means further comprises skew angle detecting means (e.g., the pump swash plate angle sensor 206 and the motor swash plate angle sensor 207 in the embodiment) which detects the tilted angle of the swash plate (e.g., the skew angle α of the pump swash plate and the skew angle β of the motor swash plate) and rotational speed detecting means (e.g., the pump rotational speed sensor 215 and the motor rotational speed sensor 216 in the embodiment) which detects the rotational speed of at least either the hydraulic pump or the hydraulic motor. Thereby, the detected moment force may be corrected on the basis of the type of the valve plate (e.g., the valve plate 150 or 160 in the embodiment) of the above mentioned at least either of the hydraulic pump or the hydraulic motor, of the skew angles of the swash plates detected by the skew angle detecting means, of the rotational speeds detected by the rotational speed detecting means, and of the operational condition of the transmission. In this way, the precision of the above mentioned adjustment of the hydraulic pressure can be improved. The term “operational condition of the transmission” refers to whether the driving force input into the transmission comes through the hydraulic pump (i.e., acceleration) or through the hydraulic motor (i.e., deceleration).

Further scope of applicability of the present invention will become apparent from the detailed description given hereinafter. However, it should be understood that the detailed description and specific examples, while indicating preferred embodiments of the invention, are given by way of illustration only, since various changes and modifications within the spirit and scope of the invention will become apparent to those skilled in the art from this detailed description.

BRIEF DESCRIPTION OF THE DRAWINGS

The present invention will become more fully understood from the detailed description given herein below and the accompanying drawings which are given by way of illustration only and thus are not limitative of the present invention and wherein:

FIG. 1 is a schematic view showing the construction of a hydraulic continuously variable transmission, which is mounted in a vehicle. For this transmission, a hydraulic controller is provided in accordance with the present invention;

FIG. 2 is a graph showing relations between the skew angles of the swash plates of the hydraulic pump and the hydraulic motor which constitute the transmission and the total speed ratio of the transmission;

FIG. 3 is a hydraulic circuit diagram showing the closed hydraulic circuit of the transmission and a circuit for providing a control pressure;

FIG. 4 is a schematic, sectional view of a regulator valve, which is used in the control pressure circuit;

FIG. 5 is a schematic, sectional view of a modulator valve, which is used in the control pressure circuit;

FIG. 6 is a graph showing a relation between the control current (I) and the control pressure (Pc) of a first linear solenoid valve;

FIG. 7 is a graph showing a relation between the control pressure (P) and the line pressure (PL) of the regulator valve;

FIG. 8 is a schematic, sectional view of a shuttle valve, which is used in the control pressure circuit;

FIG. 9 is a schematic, sectional view of a high-pressure relief valve, which is used in the control pressure circuit;

FIG. 10 is a graph showing relations between the target relief pressure (Pha and Phb) of the high-pressure relief valve and the control current (I) of second and third linear solenoid valves;

FIG. 11 is a schematic view showing the surface configuration of the (pre-compression and pre-expansion type) valve plate of the hydraulic pump and the arrangement of variable notch valves;

FIG. 12 is a schematic, sectional view of the variable notch valve;

FIG. 13 is a hydraulic circuit diagram showing a circuit which controls the skewing angles of the pump and the motor;

FIG. 14 is a schematic diagram of a control device which controls the operation of the hydraulic continuously variable transmission;

FIG. 15 is a main flow chart showing the control flow of the control device;

FIG. 16 is a graph showing relations between the vehicle speed and the engine throttle (i.e., accelerator opening) and the target rotational speed of the engine;

FIG. 17 is a subflow chart showing control determinations, which are part of the control flow shown in the main flow chart;

FIG. 18 is a subflow chart showing angular determinations, which are part of the control flow shown in the main flow chart;

FIG. 19 is another subflow chart, which shows angular determinations made in the control flow shown in the main flow chart;

FIG. 20 is another subflow chart, which shows angular determinations made in the control flow shown in the main flow chart;

FIG. 21 is a side view of a creep lever and an EXC switch, which are provided at the driver seat;

FIG. 22 is a graph showing a relation between the amount of displacement of the creep lever and the target skew angle of the swash plate of the pump;

FIG. 23 is a list showing the initial skew angles of the swash plates of the pump and the motor;

FIG. 24 is a subflow chart showing the calculation of the target skew angles of the swash plates (Step D12);

FIG. 25 is a graph showing a relation between angular deviation and the current applied to a fourth linear solenoid valve to clear target deviation;

FIG. 26 is a subflow chart, which depicts Ne control flow in the control flow shown in the main flow chart;

FIG. 27 is a graph showing relations between the skew angles of the swash plates of the pump and the motor of the hydraulic transmission and the total speed ratios, and this graph also depicts control area;

FIG. 28 is a subflow chart showing control determinations for the control area, which is part of the Ne control subflow shown in FIG. 26;

FIG. 29 is a subflow chart showing control area determinations, which is part of the subflow shown in FIG. 28;

FIG. 30 is a graph showing a relation between the speed gain coefficient of speed change which is determined in the Ne control subflow shown in FIG. 26 and the vehicle speed;

FIG. 31 is a subflow chart showing the control of the swash plate of the pump, which is part of the Ne control subflow shown in FIG. 26;

FIG. 32 is a graph showing a relation between the current inducing the target skew angle of the swash plate of the pump and the speed of speed change;

FIG. 33 is a graph showing a relation between the target current (ICMDp) applied to the fourth linear solenoid valve and the control pressure PC;

FIG. 34 is a subflow chart showing the control of the swash plate of the motor, which is part of the Ne control subflow shown in FIG. 26;

FIG. 35 is a graph showing a relation between the current inducing the target skew angle of the swash plate of the motor and the speed of speed change;

FIG. 36 is a graph showing a relation between the target current (ICMDm) applied to a fifth linear solenoid valve and a control pressure (PCM);

FIG. 37 is a subflow chart showing a lock-up control, which is part of the Ne control subflow shown in FIG. 26;

FIG. 38 is a graph showing a relation between the current inducing the lock up and the speed of speed change;

FIG. 39 is a graph showing a relation between the target current applied for the lock up and the control pressure LC for the lock up;

FIG. 40 is a subflow chart showing a relief control process, which is part of the control flow shown in the main flow chart;

FIG. 41 is a subflow chart showing a relief control process, which is part of the control flow shown in the main flow chart;

FIG. 42 is a list showing the pressures of a first and second lines which correspond to the drive direction and the acceleration or deceleration of the vehicle;

FIG. 43 is a graph showing relations between the skew angle α of the swash plate of the pump and the target relief pressure (Pha) while the vehicle is in acceleration;

FIG. 44 is a graph showing a relation between the vehicle speed V and the target relief pressure (Phb) while the vehicle is in deceleration;

FIG. 45 is a list showing the settings for the target relief pressures (Pha and Phb) under a specific condition;

FIG. 46 is a graph showing a relation between the target relief pressure (Pha or Phb) and the target control current applied to the second and third linear solenoid valves to achieve this pressure;

FIG. 47 is a subflow chart that shows variable notch control, which is part of the control flow shown in the main flow chart;

FIG. 48 is a list showing relations among the discharge of the hydraulic pump and the high pressure generated and the open or close control of the notch valve;

FIG. 49 is a graph showing a relation between the skew angle of the swash plate and the discharge of the hydraulic pump;

FIG. 50 is a skeleton diagram showing the power transmission system of the vehicle, which is equipped with the hydraulic continuously variable transmission;

FIG. 51 is a subflow chart showing a regulator control process, which is part of the control flow shown in the main flow chart;

FIG. 52 is a graph showing a relation between the target current (ICMDr) of the first linear solenoid valve and the pressure (Ph) of an oil passage in higher pressure;

FIG. 53 is a graph showing a relation between the line pressure PL and the pressure (Ph) of an oil passage at a higher pressure;

FIG. 54 is a subflow chart showing another regulator control process, which can be part of the control flow shown in the main flow chart;

FIG. 55 is a graph showing a relation between a basic target current (ICMI)r0) used for the calculation of the target current (ICMDr) of the first linear solenoid valve and the pressure (Ph) of an oil passage at a higher pressure;

FIG. 56 is a graph showing a relation between a correction factor L1 used for the calculation of the target current (ICMDr) of the first linear solenoid valve and the skew angle a of the pump swash plate;

FIG. 57 is a graph showing a relation between a correction factor L2 used for the calculation of the target current (ICMDr) of the first linear solenoid valve and the skew angle β of the motor swash plate;

FIG. 58 is a graph showing a relation between a correction factor L3 used for the calculation of the target current (ICMDr) of the first linear solenoid valve and the rotational speed Np of the hydraulic pump;

FIG. 59 is a graph showing a relation between a correction factor L4 used for the calculation of the target current (ICMDr) of the first linear solenoid valve and the rotational speed Nm of the hydraulic motor;

FIG. 60 is a graph showing a relation between the target current (ICMDr) of the first linear solenoid valve and the line pressure PL; and

FIG. 61 is a schematic view showing the surface configuration of the (lead angle type) valve plate of the hydraulic pump and the arrangement of variable notch valves.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

In the following section, a preferred embodiment of hydraulic controller used for a variable capacity hydraulic transmission is described according to the present invention. The hydraulic transmission exemplified herein comprises a hydraulic pump and a hydraulic motor, both of which are variable capacity type.

Construction of the Transmission

FIG. 1 shows a continuously variable transmission T. As seen from the figure, this transmission T can be used either for front engine front wheel drive (FF) or for rear engine rear wheel drive (RR), and it is a hydromechanical transmission, which comprises a mechanical unit 1 for drive transmission and a hydrostatic unit 2 for continuous speed variation. An engine E to drive this transmission is provided opposite the hydrostatic unit 2, making the mechanical unit 1 being positioned therebetween.

The mechanical drive transmission unit 1 comprises a power distribution mechanism 3, a power transmission mechanism 4, and a final speed reduction mechanism 5 in a first casing 1 c. The power distribution mechanism 3 comprises the input shaft 9 of the transmission, which is connected with the output shaft 7 of the engine E through a torque damper 8; a carrier 11, which is directly connected to the input shaft 9; and a pump input shaft 10, which faces the carrier 11 and extends coaxially and opposite to the input shaft 9. In addition, a plurality of pinion spindles 12 are integrally provided on the carrier 11 in such a way that the pinion spindles revolve around the input shaft 10. On each of these pinion spindles 12, a pair of large and small pinions 13 and 14 which are connected to each other in integration are pivotally provided. A small sun gear 15 with a relatively small diameter which meshes with the large pinions 13 having a relatively large diameter is provided on the input shaft 10 of the pump. In addition, a large sun gear 16 with a relatively large diameter which meshes with the small pinions 14 having a relatively small diameter is provided pivotally on the input shaft 10.

An interdrive gear 18 is provided pivotally on the pump input shaft 10 and integrally connected with the large sun gear 16, and an interdriven gear 19, which meshes with the interdrive gear 18, is provided on the output shaft 17 of the motor. These drive and driven gears 18 and 19 constitute the above mentioned power transmission mechanism 4. Furthermore, a final drive gear 20 is provided on the motor output shaft 17, and a final driven gear 21 which incorporates a differential mechanism 22 meshes with the final drive gear 20. These final drive and driven gears constitute the above mentioned final speed reduction mechanism 5. From the differential mechanism 22, a right axle 23R and a left axle 23L extend in the respective opposing directions, and through these axles 23R and 23L, the mechanical power is transmitted to drive the right and the left wheels (not shown).

The hydrostatic unit 2 for continuous speed variation comprises a hydraulic plunger type swash plate pump 24 for variable delivery, a hydraulic plunger type swash plate motor 25 for variable displacement, and a control block 27, which incorporates a hydraulic closed circuit 26 interconnecting the pump 24 and the motor 25. The control block 27 is provided between the mechanical drive transmission unit 1 and the pump 24 and the motor 25, being attached to a side of the mechanical drive transmission unit 1, such that the control block 27 pivotally supports the pump input shaft 10 and the motor output shaft 17.

The hydraulic pump 24 comprises a pump cylinder block 28 having a plurality of cylinder bores 29 around the rotational axis thereof, a plurality of pump plungers 30, which are provided slidably in the cylinder bores 29, and a pump swash plate 32 with variable skew, which slidably abuts on the shoes 31 that are pivotally attached to the tops of the plunger 30. The pump cylinder block 28 is coaxially connected to the pump input shaft 10 and in rotatably contact with a valve plate face 27 a of the control block 27.

The pump swash plate 32 is swiveled about a trunnion 33, which is provided orthogonal to the pump input shaft 10 (i.e., perpendicular to the plane of FIG. 1). This swash plate can be oriented at any position between an upright position, where the swash plate is perpendicular to the pump input shaft 10 as shown in real line in the figure (the skew angle of the pump swash plate is zero α=0), and predetermined right and left maximally tilted positions (α=αR(MAX) and aF(MAX)), where the swash plate is swiveled rightward at a predetermined maximum angle or leftward at the predetermined maximum angle as shown in dotted line in the figure. By swiveling this swash plate, it is possible to adjust the stroke of the pump plungers 30, which reciprocate in the cylinder bores when the pump cylinder block 28 is rotated. While the pump input shaft 10 is rotated by the engine E, if the swash plate is at the upright position, then the amount of stroke is zero, and the pump discharges no oil. While the swash plate is being swiveled toward the maximally tilted position with an increasing skew angle, the amount of stroke gradually increases, and the pump discharges in corresponding, increased amount. The direction of the oil flow is dependent on the direction of the tilt of the swash plate either toward the right maximally tilted position or toward the left maximally tilted position. As described in detail later, this tilt of the swash plate determines the direction of the vehicle whether it moves forward or rearward.

The hydraulic motor 25 comprises a motor cylinder block 34 having a plurality of cylinder bores 35 around the rotational axis thereof, a plurality of motor plungers 36, which are provided slidably in the cylinder bores 35, and a motor swash plate 38 with variable skew, which slidably abuts on the shoes 37 that are pivotally attached to the tops of the motor plunger 36. The motor cylinder block 34 is coaxially connected to the motor output shaft 17 and in rotatably contact with a valve plate face 27 a of the control block 27. As seen, the hydraulic motor 25 is a variable displacement motor with plungers and a swash plate.

The motor swash plate 38 is swiveled about a trunnion 39, which is provided orthogonal to the motor output shaft 17 (i.e., perpendicular to the plane of FIG. 1). The swash plate can be oriented at any position between an upright position, where the swash plate is perpendicular to the motor output shaft 17 as shown in dotted line in the figure (the skew angle of the motor swash plate is zero β=0), and a predetermined maximally tilted position (β=β(MAX)), where the swash plate is swiveled rightward at a predetermined maximum angle as shown in real line in the figure. By swiveling this swash plate, it is possible to adjust the rotation of the motor cylinder block 34 while a predetermined hydraulic pressure is applied by the pump 24 (with a constant flow of oil). When the swash plate is at the upright position, the motor cylinder block 34 is retained stationary. As the swash plate is swiveled toward the maximally tilted position with an increasing skew angle, the rotational speed of the motor cylinder block 34 gradually increases. The motor swash plate 38 can be swiveled only between the upright position and the clockwise tilted position as in the figure.

A second casing 2 c accommodates the pump 24 and the motor 25, and the second casing is combined to the control block 27, which is combined to

the first casing accommodating the mechanical unit 1.

Operation of the Transmission

Now, the operation of the above continuously variable transmission is described. When the engine E is started, the engine output is transmitted from the output shaft 7 of the engine through the torque damper 8 to the transmission. In this condition, the input shaft 9 and the carrier 11 of the transmission are driven at the same rotational speed as the output shaft 7 of the engine. As the carrier 11 is driven, the power from the engine is divided into two portions and transmitted through the large and small pinions 13 and 14 to the small and large sun gears 15 and 16.

With reference to FIG. 2, which shows relations between the skew angles α and β of the swash plates of the pump 24 and the motor 25 and the total speed ratio e of the transmission, the above mentioned division of the power is explained. The total speed ratio e is the ratio of the output rotational speed to the input rotational speed of the transmission T, and it is given in Equation (1). In FIG. 2, the ordinate represents the skew angles of the swash plates of the pump and the motor. The rightward tilts of the swash plates are plotted in the positive quadrants, and the leftward tilts are plotted in the negative quadrants. The abscissa represents the total speed ratio e, and the speed ratio in forward drive is plotted in the positive quadrants, and the speed ratio in rearward drive is plotted in the negative quadrants. In the figure, the skew angle of the swash plate of the pump is shown in real line, and that of

the motor is shown in dotted line.

Total speed ration e=(No)/(Ni)  (1)

where Ni is the rotational speed of the input shaft 9 of the transmission, and No is the rotational speed of the final driven gear 21.

When the pump swash plate 32 is at the upright position (α=0) and the motor swash plate 38 is at the maximally tilted position (β=β) (MAX)), the pump cylinder block 28 is rotatable freely, and the pump discharges no oil. The motor cylinder block 34 is hydraulically locked and retained stationary because no oil is supplied there from the pump 24. As a result, the large sun gear 16 and the interdrive gear 18 are stationary, and the small sun gear 15 rotates freely (together with the pump input shaft 10 and the pump cylinder block 28, which are connected to the small sun gear) as the carrier 11 rotates. In this condition, the engine output is wasted in idling, and no power is transmitted to the right and left axles 23R and 23L. This condition is depicted by vertical line a in FIG. 2. In this condition, the total speed ratio is zero (e=0), and the transmission T has an infinite speed change ratio.

This condition is selected when the shift lever, which is operated by the driver at the driver seat, is set at the D or R range, which is a vehicle drive range. If the shift lever is set at the P or N range, then the skew angle of the motor swash plate is controlled to become zero (β=0). As a result, the motor cylinder block 34 also becomes freely rotatable, and a neutral condition is established.

When the pump swash plate 32 is swiveled clockwise from this condition, the discharge of oil from the pump 24 starts and increases in correspondence with the increase of the skew angle of the swash plate. The oil discharged from the pump is supplied to the motor 25 to drive the output shaft 17 of the motor 25 (and the motor cylinder block 34). In this condition, if this rotational drive power is transmitted through the axles 23R and 23L to the wheels, then the wheels are driven in the forward direction. The rotational speed of the motor output shaft 17 increases as the skew angle α of the pump swash plate increases. When the skew angle becomes the maximum skew angle in the forward drive direction α F(MAX), the condition becomes the one indicated by vertical line b in FIG. 2. Therefore, the total speed ratio e increases from zero (vertical line a) to e1 (vertical line b). On the other hand, the rotational speed of the pump input shaft 10 decreases while the rotational speed of the motor output shaft 17 increases because the mechanical power transmission through the small pinion 14, the large sun gear 16, the interdrive gear 18 and the interdriven gear 19 (i.e., through the power transmission mechanism) progresses at the same time.

After the pump swash plate achieves the maximum skew angle in the forward drive direction α F(MAX) (i.e., reaches the condition indicated by vertical line b), the motor swash plate is swiveled to make the skew angle β thereof gradually small from the maximum angle. As the skew angle becomes smaller and smaller, the rotational speed of the motor output shaft 17 increases further above the speed which is indicated by vertical line b. When the skew angle β becomes zero (i.e., the swash plate comes to the upright position), the rotational speed of the motor output shaft reaches the maximum speed (i.e., the total speed ratio becomes e2, and this condition is indicated by vertical line c in the figure).

However, as mentioned above, while the rotational speed of the motor output shaft 17 increases, the mechanical power transmission through the power transmission mechanism 4 increases, and the rotational speed of the pump input shaft 10 decreases. The gear ratio of the power distribution mechanism 3 and the power transmission mechanism 4 is predetermined in such a way that the rotational speed of the pump input shaft 10 (and the pump cylinder block 28) becomes zero when the skew angle β of the motor swash plate becomes zero (i.e., the swash plate comes to the upright position). Therefore, when the motor swash plate is at the upright position (β=0), the motor cylinder block 34 is freely rotatable, and the pump cylinder block 28 is hydraulically locked and becomes stationary. In this condition (which is indicated by vertical line c in the figure), only the mechanical power transmission through the power transmission mechanism 4 is performed.

On the other hand, if the pump swash plate 32 is swiveled counterclockwise, starting from the condition indicated by vertical line a in the figure, then the oil is discharged from the pump 24 in the direction opposite to that described above in the hydraulic closed circuit 26. This reversed flow of oil to the motor 25 drives the motor output shaft 17 (and the motor cylinder block 34) in the direction opposite to that described above (i.e., in the reverse drive direction). The rotational speed of the motor output shaft 17 increases as the skew angle α of the pump swash plate increases. When the skew angle becomes α R(MAX), the condition reaches the point which is indicated by vertical line d in FIG. 2. In this way, the total speed ratio e changes from zero (vertical line a) to a negative value (e3).

Hydraulic Closed Circuit for Power Transmission

Now, with reference to FIG. 3, the hydraulic closed circuit 26 and the hydraulic circuit control system of the hydrostatic unit 2 for continuous speed variation are described. In this figure, the hydraulic pump 24 and the hydraulic motor 25 are indicated in symbols. The hydraulic closed circuit 26 comprises a first oil passage 26 a which connects one port 24 b of the pump 24 with one port 25 a of the motor 25 and a second oil passage 26 b which connects the other port 24 a of the pump 24 with the other port 25 b of the motor 25.

As mentioned previously, the swash plate 32 of the pump 24 is swiveled in either direction, clockwise or counterclockwise, from the upright position (neutral position). When the pump swash plate is swiveled clockwise (i.e., in the forward drive direction), the oil sucked from the port 24 a is discharged from the port 24 b. This oil is supplied to the motor 25 through the port 25 a to drive the motor 25 in the forward drive direction. Then, the oil is discharged from the port 25 b, again to be sucked into the port 24 a, thus circulating in the closed circuit 26. In this condition, if the wheels are driven by the rotation of the motor 25, then the pressure in the first oil passage 26 a becomes high in correspondence, and the pressure in the second oil passage 26 b become low. On the other hand, if the rotational speed of the wheels are being reduced by engine brake in coasting, for example, then the pressure in the second oil passage 26 b becomes high in proportion to the force of engine brake, and the pressure in the first oil passage 26 a become low.

When the pump swash plate 32 is swiveled counterclockwise (i.e., in the rearward drive direction), the oil flow is reversed from that described above. This reversed oil flow drives the motor in the rearward drive direction. Likewise, the conditions of the pressure in the first and second oil passages 26 a and 26 b become opposite of what is described above.

In this way, the power transmission between the hydraulic pump and the hydraulic motor is performed. However, in this power transmission, while the oil is circulated in the hydraulic closed circuit 26, heat is generated, and the temperature of the oil rises. Also, in the circulation, contaminants may be collected in the oil, or some part of the oil may leak, for example, through the clearance of the plungers, into an oil tank. To solve such problems, the system enables part of the oil to be exchanged for cooling, supplementary feed, and purification (or flashing). A charge pump 43 is provided to supply oil in the oil tank 41 to a first line 100 through a suction filter 42. The charge pump 43 is driven by the engine E directly, so the amount of discharge is proportional to the rotation of the engine.

The oil discharged from the charge pump 43 into the first line 100 is regulated by a regulator valve 60 to have a predetermined line pressure PL. The first line 100 is branched as shown in the figure, and one branched first line 100 a is connected to a modulator valve 65, which includes a reducing valve. The output port of the modulator valve 65 is connected to a second line 101 to adjust the pressure of the second line to a predetermined modulated pressure Pm FIG. 5 shows the construction of the modulator valve 65, which includes a spool 66, which is biased leftward by a spring 67, in a housing. The hydraulic pressure in the branched first line 100 a, which is connected to a port 65 a of the modulator valve 65, is reduced to a pressure (a constant pressure)at which the force of the spring 67 balances with the force of the pressure in the control line 101, thus creating the modulated pressure Pm in the control line 101. In the figure, mark “x” indicates a drainage line.

The second line 101 also branches into a plurality of lines, and one of the branched line, a branched second line 101 a, leads to a first linear solenoid valve 51. The first linear solenoid valve 51 adjusts the modulated pressure Pm to generate a control pressure PCL in proportion to a control current (I) as shown in FIG. 6, and the control pressure PCL acts on the regulator valve 60 through a control line 110.

FIG. 4 shows the construction of the regulator valve 60, which comprises a spool 61, which is slidable leftward and rightward, and a spring 62, which biases the spool 61 leftward, in a housing. The regulator valve has a plurality of ports 60 a-60 e in the housing. The ports 60 a and 60 b are connected to the first line 100, the port 60 c is connected to a charge line 130, the port 60 d is connected to a discharge line 131, and the port 60 e is connected to the above mentioned control line 110.

In this condition,the left end of the spool 61 of the regulator valve 60 is exposed to the line pressure PL from the first line 100 through the internal fluid- communication hole 61 a, and the right end is pushed by the biasing force of the spring 62 and exposed to the control pressure PCL. As mentioned above, the control pressure PCL is adjustable with the first linear solenoid valve 51, so the line pressure PL is controllable by controlling the control current (I), which is applied to the first linear solenoid valve 51, as shown in FIG. 7. In the control of the line pressure PL with the regulator valve 60, excess oil flows into the charge line 130 when the port 60 a comes into fluid communication with the port 60 c as the spool 61 shifts to the right. Further excess oil flows into the discharge line 131 when the port 60 b comes into fluid communication with the port 60 d.

As shown in FIG. 3, the charge line 130 is connected with a charge supply lines 105 a and 106 a, which are connected with the first and second oil passages 26 a and 26 b respectively through check valves 44 a and 44 b. The oil flowing into the charge line 130 must pass through either the check valve 44 a or the check valve 44 b before being supplied into either the first oil passage 26 a or the second oil passage 26 b whose pressure is lower than the other. In this way, the oil in the hydraulic closed circuit 26 is replenished.

The oil discharged into the discharge line 131 is cooled by an oil cooler 151, and then it is returned through a lubricator 152 into the tank 41.

As shown in FIG. 3, the first and second oil passage 26 a and 26 b constituting the hydraulic closed circuit 26 are connected with discharge lines 105 b and 106 b respectively, and these discharge lines 105 b and 106 b are connected to a shuttle valve 70. FIG. 8 shows the construction of the shuttle valve 70, which comprises a spool 71, which is slidable rightward and leftward in a housing, a pair of springs 72 and 73, one of which biases the spool 71 to the right, and the other to the left.

The hydraulic pressures of the discharge lines 105 b and 106 b work on the right and left ends of the spool 71 respectively. When the pressure of either one of the first or second oil passage 26 a or 26 b becomes higher than the other, the spool 71 is shifted by the pressure difference, and either the left side port 70 a or the right side port 70 b of the shuttle valve 70 whose hydraulic pressure is lower than the other is connected with the port 70 c which leads to a discharge line 132 to discharge some oil from the line which has the lower pressure. In this way, the oil is discharged in the amount which corresponds with the amount replenished in the hydraulic closed circuit 26, and the hydraulic oil is cooled and flashed, etc. In addition, the discharge line 132 is provided with a low pressure relief valve 74, which regulates the pressure of the line having the lower pressure. The oil discharged into the discharge line 132 is also cooled by the oil cooler 151 and returned through the lubricator 152 into the tank 41.

The hydraulic closed circuit 26 is also provided with high pressure relief valves 75F and 75R, which are connected through relief lines 105 c and 106 c to controls the maximum pressure of the first oil passage 26 a and the second oil passage 26 b. These relief valves are identical in construction, so here, the high relief valve 75F, one of these two valves, is described with reference to FIG. 9.

The high relief valve 75F has two independent spools 76 and 77 in a housing. A first spool 76 is biased rightward by a first spring 79 a to cut off the fluid communication between the right side port 75 a and a port 75 b located inward in the housing, but if the spool is shifted leftward against this biasing force, then the two ports 75 a and 75 b are brought into fluid communication.

The port 75 a is connected to the first oil passage 26 a through the relief line 105 c, and the port 75 b is connected to a branched charge line 130 a. In addition, an orifice 76 a is provided in the first spool 76, so the pressure in the first oil passage 26 a provided through the port 75 a acts on both the right and left sides of the first spool 76. Therefore, the first spool 76 is positioned at the right end of its stroke by the bias of the first spring 79 a in normal condition.

On the other hand, the second spool 77 is biased leftward by a second spring 79 b whose right end is also biasing a blocking valve piece 78 rightward in the housing. In this condition, the blocking valve piece 78 blocks a fluid communication passage 79 c which is provided with an orifice to communicate with the valve chamber of the first spool 79 a. The left side port 75 c, to which the left end of the second spool 77 faces, is connected through a control line 107 to a second linear solenoid valve 52, which adjusts the line pressure PL of a branched line 100 b in correspondence with a control current to provide a control pressure PCH to the control line 107. In this condition, the pressure on the left end of the second spool 77 can be controlled by controlling the control current applied to the second linear solenoid valve 52.

In the high pressure relief valve 75F, the pressure of the first oil passage 26 a acts on the blocking valve piece 78 from the right, and the force of the second spring 79 b act on the blocking valve piece 78 from the left. Because the left end of the second spring 79 b is placed on the second spool 77, the force of the control pressure PCH from the control line 107 acting on the second spool 77 is added to the rightward biasing force. Therefore, the biasing force acting on the blocking valve piece 78 rightward is controllable by adjusting the control pressure PCH, which is controlled by the second linear solenoid valve 52.

By this rightward biasing force, the blocking valve piece 78 blocks a fluid communication passage 79 c. However, if the pressure of the first oil passage 26 a rises, and the force from this pressure acting on the blocking valve piece 78 leftward exceeds the above rightward biasing force, then the blocking valve piece 78 is shifted leftward opening the fluid communication passage 79 c to the drainage. As a result, an oil flow is generated through the orifice 76 a of the first spool 76, which in turn creates a difference between the pressures on the right and left sides of the first spool 76. By this pressure difference, the first spool 76 is shifted leftward, and the port 75 a and the port 75 b are brought into fluid communication. As a result, some oil is discharged from the first oil passage 26 a to the branched charge line 130 a, and this discharged oil is fed into the second oil passage 26 b through the check valve 44 b.

In summary, when the pressure of the first oil passage 26 a rises above a predetermined value, the blocking valve piece 78 is released, and the first spool 76 shifts to the left, which discharges oil from the first oil passage 26 a to the second oil passage 26 b to maintain the pressure of the first oil passage 26 a at or below the predetermined value. On the other hand, if the pressure of the second oil passage 26 b rises above the predetermined value, the other high pressure relief valve 75R discharges oil from the second oil passage 26 b to the first oil passage 26 a to prevent the pressure of the second oil passage 26 b from rising above the predetermined value. In this way, the pressures of the first and second oil passages 26 a and 26 b are prevented from rising above the predetermined value by the high pressure relief valves 75F and 75R, which value or pressure is adjustable by controlling the current applied to the second and third linear solenoid valves 52 and 53 as described above. In this example, the high relief pressure (PHF and PHR) is adjustably set in proportion with the control current (I) as shown in FIG. 10.

In this embodiment, the control block 27 of the hydraulic pump 24 and the motor 25 is provided with respective valve plates, each having a configuration shown in FIG. 11, in the valve plate face 27 a. These valve plates, one provided for the pump and the other for the motor, may differ from each other in size, but their basic configurations and functions are identical. Thus, only the valve plate 150 of the pump 24 is described here with reference to FIG. 11.

In the pump 24, the pump cylinder block 28 is driven by the engine and is rotated in the direction indicated with arrow A in the figure (clockwise) in contact with the valve plate 150. When the pump swash plate 32 is swiveled in the forward drive direction, the pump plungers 30 reciprocate in the cylinder bores in synchronization with the rotation of the pump cylinder block 28. In this reciprocation, the pump plungers 30 reach the top dead center (T.D.C.) at the lower end of the valve plate as shown in FIG. 11, and they reach the bottom dead center (B.D.C.) at the upper end of the valve plate as the cylinder block rotates. While the pump plungers 30 travel from the top dead center (T.D.C.) to the bottom dead center (B.D.C.) above the left half of the valve plate as shown in the figure, oil is sucked, and while they travel from the bottom dead center (B.D.C.) to the top dead center (T.D.C.)above the right half of the valve plate, the oil is discharged.

Therefore, the valve plate 150 is provided with a semicircular first port 151 in the left half and with a semicircular second port 152 in the right half of the valve plate. The first port 151 is connected to the second oil passage 26 b, and the second port 152 is connected to the first oil passage 26 a. In addition, the valve plate is provided with main notches 151 a and 152 a, which extend to the respective entrances of the first and second ports 151 and 152 in the rotational direction. These main notches are to moderate the rapid pressure change which occurs when the cylinder bores 29 are brought into fluid communication with the ports 151 and 152 as the pump cylinder block 28 rotates. Furthermore, sub notches 153 and 154 are provided independently from and in parallel with the main notches 151 a and 152 a as shown in the figure.

As the sub notches 153 and 154 are independent and away from the main notches 151 a and 152 a and the ports 151 and 152, no notch effect can be expected in this condition. However, the sub notches 153 and 154 are connected to the inlets of the first and second ports 151 and 152 through shortcircuiting oil passages 155 a and 155 b and 156 a and 156 b as shown in dotted line in the figure, and variable notch valves 80A and 80B are provided on the shortcircuiting oil passages to open and close the shortcircuiting oil passages.

Both the variable notch valves 80A and 80B have an identical construction, which is shown in the FIG. 12 (this figure shows the variable notch valve 80A). The variable notch valve 80A comprises a valve spool 81, which is slidable rightward and leftward, a spring 82, which biases the valve spool 81 rightward, and a supporting spool 83, which slidably engages with the left end portion of the valve spool 81, in a housing. The housing is provided with a first port 80 a, which is connected to the shortcircuiting oil passage 155 a, a second port 80 b, which is connected to the shortcircuiting oil passage 155 b, and a third port 80 c, which is connected to a control line 102. When the valve spool 81 is biased by the force of the spring 82 and positioned at the right side as shown in the figure, the right end portion of the spool 81 cuts off the fluid communication between the first and second ports 80 a and 80 b. When the valve spool 81 is shifted to the left, theses ports are brought into fluid communication. The pressure in the first port 80 a is communicated through a pore 81 a which is provided in the spool 81 to the chamber 84 which exists between the valve spool 81 and the supporting spool 83. Therefore, this pressure will not provide any thrust to the valve spool 81.

As shown in FIG. 3, the control line 102 is connected with a branched second line 101 b through an orifice 45 a and to an open- close control solenoid valve 45, which can open the control line 102 to the drainage. When the open-close control solenoid valve 45 opens the control line 102 to the drainage, the pressure in the control line 102 is lowered. On the other hand, when the open-close control solenoid valve 45 closes or cuts off the fluid communication of the control line 102 to the drainage, the modulated pressure Pm is supplied from the branched second line 101 b to the control line 102.

When the control line 102 is maintained at the low pressure by the opening of the open-close control solenoid valve 45, the valve spool 81 of the variable notch valve 80A is positioned at the right side by the biasing force of the spring 82, blocking the shortcircuiting oil passages 155 a and 155 b. In this condition, the sub notch 153 has no effect on the system. On the other hand, when the control line 102 is supplied with the modulated pressure Pm by the closing of the open-close control solenoid valve 45, the valve spool 81 is shifted to the left side by this pressure, bringing the shortcircuiting oil passages 155 a and 155 b into fluid communication. In this condition, the sub notch 153 becomes effective.

This description is made of the variable notch valve 80A, but the same can be said of the other variable notch valve 80B, which is controllable in the same way as the variable notch valve 80A. Also, the hydraulic motor 25 is provided with such sub notches, and the pressure of a control line 103 is variably set with an open-close control solenoid valve 46 to control variable notch valves 80C and 80D, which control the sub notches.

Skew Angle Control System for the Swash Plates

Now, a control system for swiveling the pump swash plate 32 and the motor swash plate 38 is described with reference to FIG. 13. This control system uses the line pressure PL, which is supplied through a branched first line 100 d, and the modulated pressure Pm, which is supplied through a branched second line 101 c. In FIGS. 3 and 13, circled letters “a” and “b” in one of the figures are connected to the respective circled letters of the other figure.

For swiveling the pump swash plate 32, a pair of servo cylinders 92 a and 92 b are provided slidably in servo cylinder bores 91 a and 91 b as shown in the figure, and the servo cylinder bores 91 a and 91 b are connected through servo control lines 121 and 122 to a pump control valve 84, which is a four way valve. Depending on the position of the spool 85, the pump control valve 84 switches the supply of the line pressure PL from the branched first line 100 d either to the servo control line 121 or to the servo control line 122. The spool 85 is biased rightward by a spring 86, and the right end of the spool receives the pressure coming through the right end port 84 a. The balance of the forces from the spring and the pressure sets the position of the spool in the housing.

When the pressure of the right end port 84 a is controlled, the position of the spool 85 is adjusted to operate the servo cylinders 92 a and 92 b, which eventually regulates the pump swash plate 32. To control the pressure of the right end port 84 a, a pump control pressure PCP is supplied to the port 84 a from a fourth linear solenoid valve 54 through a control line 111. The fourth linear solenoid valve 54 controls the modulated pressure Pm of the branched second line 101 c to generate the pump control pressure PCP in proportion with a control current and leads the pump control pressure into the control line 111. Thus, by controlling the control current applied to the fourth linear solenoid valve 54, the pump control valve 84 is controlled to control the skew angle of the pump swash plate 32.

The control of the skew angle of the motor swash plate 38 is operated in the same manner. A pair of servo cylinders 96 a and 96 b are provided slidably in servo cylinder bores 95 a and 95 b, which are connected through servo control lines 123 and 124 to a motor control valve 87. In the same way as the pump control valve 84, the pressure of the right end port 87 a is controlled to control the position of the spool 88, which controls the servo cylinders 96 a and 96 b to regulate the motor swash plate 38. To control the pressure of the right end port, a motor control pressure PCM is supplied to the port 87 a from a fifth linear solenoid valve 55 through a control line 112. Thus, by controlling the control current applied to the fifth linear solenoid valve 55, the motor control valve 87 is controlled to control the skew angle of the motor swash plate 38.

In this embodiment, the hydraulic pump 24 is provided with a lock-up brake 93, which can hold the pump cylinder block 28 stationary. As previously mentioned, when the skew angle of the pump swash plate becomes the maximum skew angle in the forward drive direction, i.e., α=α F(MAX), and the skew angle of the motor swash plate becomes zero (β=0) in the forward drive condition (represented by vertical line c in FIG. 2), in theory, the rotational speed of the pump cylinder block 28 comes to zero, and the power transmission is carried out only through the power transmission mechanism 4 (if there is no transmission loss). However, in reality, there is a loss from frictional drag, oil leak, etc., which causes the pump cylinder block 28 to rotate a little and reduces the efficiency of the power transmission through the power transmission mechanism 4. To solve this problem, the pump cylinder block 28 is held stationary with the lock-up brake 93 to perform only the mechanical power transmission and to improve the transmission efficiency.

The lock-up brake 93 is a wet multiple disc brake, and it is operated in cooperation with a piston 104. The lock-up pressure PLB which empowers the piston 104 to squeeze the brake is generated by a sixth linear solenoid valve 56 and supplied through a lock-up line 113. The sixth linear solenoid valve 56 generates the lock-up pressure PLB, which is proportional to the control current supplied to the solenoid valve. Therefore, it is possible to control the lock-up brake 93 to engage partially or completely as much as desired.

Operational Control

Now, the operational control of the variably continuous transmission T, which has the above construction and control circuit, is described. FIG. 14 shows the general construction of the control system. A control unit ECU generates control signals to the solenoid valves 45, 46 and 51-56 (these valves constitute the hydraulic circuit, and they have been already explained) in response to detection signals which are received from various sensors 201-214. The control unit ECU comprises an input interface to receive signals from the sensors and an output interface to send signals to the solenoid valves. The control unit ECU is operated by the central processing unit CPU, which operates on a software program that is stored in memory ROM 1 and ROM 2 to generate the control signals.

The sensors include: an engine rotation sensor 201, which detects the rotational speed Ne of the engine E; a vehicle speed sensor 202, which detects the speed of the vehicle V; a shift position sensor 203, which detects the position of the shift lever; a brake sensor 204, which detects the operation of the brake of the vehicle; a throttle opening sensor 205, which detects the degree of the throttle opening θ TH or how much the engine throttle is opened; a pump swash plate angle sensor 206, which detects the skew angle a of the pump swash plate; a motor swash plate angle sensor 207, which detects the skew angle β of the motor swash plate; a pressure sensor 208, which detects the pressure P1 of the first oil passage 26 a that rises high when an external driving force acts on the transmission through the hydraulic pump 24 (i.e., the vehicle is in acceleration); a pressure sensor 209, which detects the pressure P2 of the second oil passage 26 b that rises high when an external driving force acts on the transmission through the hydraulic motor 25 (i.e., the vehicle is in deceleration); a creep lever sensor 210, which detects the amount of manipulation of the creep lever operated by the driver to set a creeping force; an EXC switch sensor 211, which detects the operation of an EXC switch that renews the initial value for the skew angle of the pump swash plate; a PB sensor 212, which detects the negative pressure of the engine in suction; a transmission oil temperature sensor 213, which detects the temperature of the oil in the transmission; an engine water temperature sensor 214, which detects the temperature of the engine cooling water; a pump rotational speed sensor 215, which detects the rotational speed Np of the hydraulic pump; and a motor rotational speed sensor 216, which detects the rotational speed Nm of the hydraulic motor.

The operational control of the transmission follows a main control flow which is shown in FIG. 15. At first, at Step B1, an initialization control is executed to set an initial value to the output current which is supplied to the solenoid valves. Then, a target engine rotational speed Ne′ is set on the basis of the vehicle speed V, the throttle opening θ TH and the shift lever position at Step B2. This step is executed, for example, by using a map such as shown in FIG. 16. The map shows the target engine rotational speed Ne′ in ordinate in relation with the vehicle speed V in abscissa for respective throttle openings. When the vehicle speed V and the throttle opening θ TH at present are detected, a value for the target engine rotational speed Ne′ is taken in correspondence from the map. The map shown in FIG. 16 is for the condition where the shift lever is positioned at the D range, and different maps are prepared for other conditions where the shift lever is positioned at other ranges. In this way, with the detection of the shift lever position, the appropriate map is selected in correspondence with the position of the shift lever to set the target engine rotational speed Ne′.

Then, the control flow proceeds to Step B3, and a determination is made to proceed either to an angle control (B4) or to an Ne control (B5)(the details of these controls are described later). The angle control (B4) is to set the skew angle of the pump swash plate to a predetermined angle. This will be described more in detail later. The Ne control is to control the skew angles of the pump swash plate and motor swash plate and the operation of the lock-up brake in feedback on the basis of the engine rotational speed Ne. This will be also described more in detail later.

While either the angle control (B4) or the Ne control (B5) is being executed, the high pressure relief valves 75F and 75R are controlled at Step B6; the operation of the variable notch valves 80A-80D is controlled at Step B7; and the line pressure PL is set by the regulator valve 60 at Step B8.

These steps are repeated as shown in FIG. 15, and these steps will be also described more in detail later.

Now, the determination performed at Step B3 is described with reference to FIG. 17. At first, a determination is made whether the shift position is in the N or P range or not at Step C1. If the shift position is in the N or P range, then the angle control B4 is performed. If the shift position is in any other range, then another determination is made whether the vehicle speed V is in acceleration or not, i.e., whether the speed is increasing or not at Step C2. If the vehicle is not in acceleration, then yet another determination is made whether the vehicle speed V is equal to or less than a predetermined speed V1 (e.g., 5.0 km/h) or not at Step C3. If it is equal to or less than the predetermined speed, then the angle control B4 is performed. However, if it is above the predetermined speed, then the control flow proceeds to Step C4.

On the other hand, if the vehicle is determined as being accelerating at Step C2, then the control flow proceeds to Step C6, and a determination is made of the throttle opening θ TH of the engine whether the throttle opening is about zero or not (i.e., whether the accelerator pedal is released or not). If the throttle opening θ TH is determined to be not zero, then the control flow proceeds to Step C4. However, if it is determined as almost zero, then the control flow proceeds to Step C7, where another determination is made whether the vehicle speed V is equal to or less than the predetermined speed. If it is so, then the angle control B4 is performed. The driver is judged as not in consideration of accelerating the vehicle if the throttle is closed completely and the vehicle is running very slowly (at the predetermined speed or less) even though the vehicle is in acceleration, so the angle control B4 is carried out. However, if the vehicle speed V is above the predetermined speed, then the control flow proceeds to Step C4.

At Step C4, a determination is made of the actual rotational speed Ne of the engine whether the actual rotational speed Ne is less than the target engine rotational speed Ne′, which is set at Step B2. If the actual rotational speed Ne is determined to be greater or equal to the target rotational speed Ne′, then the Ne control B5 is carried out. On the other hand, if the actual rotational speed Ne is determined to be less than the target rotational speed Ne′, then another determination is made of the skew angle α of the pump swash plate whether the angle α is equal to or smaller than a predetermined value α′ or not. If it is equal to or smaller than the predetermined value, then the angle control B4 is carried out. However, if it is above the predetermined value, then the Ne control B5 is carried out. For this predetermined value α′, the control flow uses the target skew angle α′ for the swash plate that is determined in another control method determination flow, which will be described later with reference to FIGS. 18 and 19.

The above control flow for control method determination makes clear that the angle control B4 is performed in the following conditions:

(1) the position of the shift lever is at the N (neutral) or P (parking);

(2) the vehicle is not in acceleration, and it is driving at a speed which is equal to or less than the predetermined speed;

(3) the throttle is closed completely, and the vehicle speed is equal to or less than the predetermined speed even though the vehicle is in acceleration; and

(4) the skew angle α of the pump swash plate is equal to or smaller than the predetermined value, and the actual engine rotational speed Ne is less than the target engine rotational speed Ne′.

Now, the angle control B4 is described with reference to FIGS. 18 and 19. It is clear from the above mentioned control flow that the angle control B4 is to be performed when the transmission is close to its neutral condition (in the vicinity of vertical line a in FIG. 2). The skew angle β of the motor swash plate is set either at zero or at the maximum value β (MAX). This means that when the shift position is in the N or P range, the skew angle β is set at zero, and the skew angle β is set at the maximum value β (MAX) in other cases. The skew angle of the motor swash plate is easily set at these angles by adjusting the motor control pressure PCM, which is supplied to the right end port 87 a of the motor control valve 87, to zero or to the maximum value so as to shift the spool 88 to the left end position or to the right end position.

In addition to the above control of the skew angle of the motor swash plate, the angle control B4 is to carry out the positional feedback control of the skew angle of the pump swash plate. In this control, first, a determination is made whether the shift position is in the N or P range or not at Step D1. If the shift position is in the N or P range, then the target deviation a1 of the pump swash plate is set at zero (α1=0) at Step D6.

If the shift position is in any range other than the N or P range, then a determination is made of the engine rotational speed Ne whether the rotational speed Ne is equal to or less than a predetermined rotational speed (e.g., 600 rpm, which is a little less than an idling rotational speed, i.e., 700 rpm) at Step D2. If the engine rotational speed Ne is equal to or less than the predetermined rotational speed, then the target deviation α1 of the pump swash plate is set at zero (α1=0) at Step D6. From this control, it is understood that when the positional feedback control for the pump swash plate (angle control) is performed, the target deviation al is set to zero (α1=0) unconditionally if the engine rotational speed is equal to or less than the idling rotational speed so as to prohibit any control which may increase the delivery of the pump.

On the other hand, if the engine rotational speed is above the predetermined value, then a determination is made whether the engine throttle opening is completely closed (θTH=0) or not at Step D3. If the throttle is open, then the control flow proceeds to Step D17. However, if the throttle is completely closed, then the control flow proceeds to Step D4, and another determination is made whether the vehicle speed V is equal to or less than a predetermined vehicle speed (e.g., 5 km/h) or not. If the vehicle speed V is above the predetermined speed, then the control flow proceeds to Step D17. However, if it is equal to or less than the predetermined vehicle speed, then the control flow proceeds to Step D5, where another determination is made whether the brake operation is ON or not, i.e., whether the brake pedal is stepped on or not. If the brake is ON, then the control flow proceeds to Step D6. If it is OFF, then the control flow proceeds to Step D9.

Consequently, the target deviation α1 of the pump swash plate is set for each case in the following steps. First, the case where the control flow has proceeded to Step D9 is described. At Step D9, a determination is made whether the shift position is in any of the forward drive ranges (D, L, S or M range) or in the reverse drive range (R range). If it is in a forward drive range, then the control flow proceeds to Step D10. However, if the shift position is in the reverse drive range, then the control flow proceeds to Step D15.

First, a description is made for the former case, where the shift position is in a forward drive range. As shown in FIG. 21, a creep lever 220 and an EXC switch 221 are provided at the driver seat of the vehicle, in which this transmission is incorporated. The creep lever 220 is manipulated by the driver, and the EXC switch 221 is used to renew the initial value of the skew angle of the pump swash plate. At Step D10, the target deviation α1 of the pump swash plate is set on the basis of the manipulated amount ACR of the creep lever 220. Specifically, the relation between the manipulated amount ACR of the creep lever 220 and the target deviation α1 of the pump swash plate is predetermined as shown in FIG. 22, and the setting is performed in accordance with this relation.

As shown in FIG. 21, the creep lever 220 is swiveled forward and backward from an upright position, which is a neutral position N. The creep lever 220 is an automatically resetting lever which automatically returns to the neutral position, so when it is not manipulated, it is positioned at the neutral position N as shown in the figure. When the creep lever 220 is at the neutral position N, the target deviation α1 of the pump swash plate is zero. If the lever 220 is swiveled to the “+” side as shown in the figure, then the target deviation α1 is set to a positive value which is proportional to the manipulated amount of the creep lever. Likewise, if the creep lever is swiveled to the “−” side, then the target deviation α1 is set to a negative value which is proportional to the manipulated amount of the creep lever. However, this operation is effective only for the case where the shift position is determined at Step D1 as being in a range other than the N or P range and the rotational speed of the engine is determined at Step D2 as being equal to or greater than a predetermined rotational speed (e.g., 600 rpm). Therefore, if the engine rotational speed is less than the predetermined value, the target deviation α1 of the pump swash plate is set to zero so that the skew angle of the pump swash plate will not increase, notwithstanding the operation of the creep lever 220.

At Step D11, the initial value α0 of the skew angle of the pump swash plate and the initial value β0 of the skew angle of the motor swash plate are set. These values are set on the basis of the shift position as shown in FIG. 23. The conditions which have led the control flow to Step D11 are that the vehicle speed is less than the predetermined speed, that the throttle opening is zero, i.e., closed, that the brake is OFF and that the vehicle is in the D, L, S, or M range. Therefore, in this case, the initial value α0 of the skew angle of the pump swash plate is set to predetermined value 2, and the initial value b0 of the skew angle of the motor swash plate is set at the maximum (β0=MAX).

From the target deviation α1 of the pump swash plate and the initial value α0 of the skew angle, a target skew angle α′ is calculated for the pump swash plate at Step D12. This calculation is described with reference to FIG. 24. As mentioned previously, the target deviation α1 of the pump swash plate is set when the creep lever 220 is manipulated by the driver. More specifically, the target deviation α1 of the pump swash plate is taken for the calculation of the target skew angle α′ when the driver presses the EXC switch 221 down while he is manipulating the creep lever (i.e., when the driver shows his intention to set the value). The EXC switch 221 is a momentary switch, which is ON only while it is pressed down, and it returns to OFF when released.

At this step, a determination is made at Step D51 whether the flag which is set up when the EXC switch is turned on is one (EXC switch=1) or not. If the flag is zero (EXC switch=0), then the control flow proceeds to Step D54, and another determination is made of the target deviation α1 of the pump swash plate whether the target deviation is zero (α1=0) or not. If the target deviation is not zero, i.e., the creep lever 220 is being manipulated, then the control flow proceeds to Step D55, where the target skew angle α′ is calculated (α′=α0+α1). At this point, a timer is set to count a predetermined time period at Step D56, and a determination is made of the condition of the EXC switch 221 whether the EXC switch is turned on or not at Step D57. If the switch is OFF, then the control flow return to the start of this control flow. However, if the switch is turned ON, then the flag is set up (EXC switch=1) at Step D58, and the target skew angle α′ at this moment is memorized as α ex at Step D59.

When the flag is one (EXC switch=1), the control flow returns to the start of this control flow and proceeds to Step D51 and to Step D52, and the target skew angle (α′=α ex) is retained until the time which is set on the timer elapses at Step D53. When the time is up, a determination is made of the target deviation α1 of the pump swash plate whether the target deviation is zero (α1=0) or not at Step D54. If it is zero (α1=0), i.e., the creep lever 220 is released after the EXC switch 221 is pressed, the control flow proceeds to Step D60. There, another determination is made whether the flag is up (EXC switch=1) or not. If the flag is up (1), then the initial value α0 for the skew angle of the pump swash plate is reset at Step D61 with the target skew angle α ex which is memorized at Step D59, and the flag is lowered (EXC switch=0) at Step D62. This reset initial value α0 is then set as the target skew angle α′ at Step D63.

Summarizing the above control at Step D12, the following can be said. If the EXC switch 221 is pressed down while the creep lever 220 is manipulated, then the target skew angle α′ of the pump swash plate is set by adding the initial value α0 of the skew angle of the pump swash plate to the target deviation α1 which corresponds with the amount of manipulation of the creep lever at the moment. Then, this target skew angle α′ is retained even though the creep lever 220 is returned to the neutral position. In addition, this target skew angle α′ (=α ex) becomes a new initial value α0 for the skew angle of the pump swash plate. Thereafter, if the EXC switch 221 is pressed down while the creep lever 220 is manipulated again, then the current target skew angle α′ of the pump swash plate is added with the target deviation α1 which corresponds with the amount of manipulation of the creep lever made at this moment. Thus, if the manipulation of the creep lever 220 and the pressing of the EXC switch 221 are repeated, then the target skew angle α′ is set accordingly with the addition of the target deviations α1 which correspond to the repeated operations.

After the target skew angle α′ of the pump swash plate is set at Step D12 as described above, a determination is made at Step D13 whether this value α′ is greater than the initial value α0, which is set at Step D11. If the target skew angle α′ is equal to or smaller than the initial value α0 (α′≦α0), then the target skew angle α′ is set equal to the initial value (α′=α0) at Step D14. When the creep lever 220 is manipulated to the negative side, the target skew angle α′ may become smaller than the initial value (α′≦α0). However, even if such a case happens, the skew angle of the pump swash plate will not set to a value smaller than the initial value α0.

On the other hand, if the shift position is in the reverse drive range (R range), then a control similar to the above is executed at Step D15. The target deviation α1 of the pump swash plate is set in correspondence with the manipulation of the creep lever 220, and then the control flow proceeding to Step D16, the initial value α0 of the skew angle of the pump swash plate and the initial value β0 of the skew angle of the motor swash plate are set to the values which are listed for the R range in FIG. 23. In this case, the vehicle speed is less than the predetermined speed, the throttle opening is zero, i.e., closed, and the brake is OFF. Thus, the initial value α0 of the skew angle of the pump swash plate is set to Predetermined value 1, and the initial value β0 of the skew angle of the motor swash plate is set to the maximum (β0=MAX).

From the target deviation α1 of the pump swash plate and the initial value α0 of the skew angle, a target skew angle α′ is calculated at Step D12 in the same way as described above. Then, a determination is made at Step D13 whether this value α′ is greater than the initial value α0, which is set at Step D16. If the target skew angle α′ is equal to or smaller than the initial value α0 (α′≦α0), then the target skew angle α′ is set equal to the initial value (α′=α0) at Step D14.

Now, a description is made for the case where the control flow has proceeded to Step D17. At Step D17, a determination is made whether the shift position is in any of the forward drive ranges (D, L, S or M range) or in the reverse drive range (R range). If it is in a forward drive range, then the control flow proceeds to Step D18. However, if the shift position is in the reverse drive range, then the control flow proceeds to Step D20.

In either case, if the shift position is in a forward drive range, then at Step D18, or if the shift position is in the reverse drive range, then at Step D20, the target deviation α1 of the pump swash plate is set in correspondence with the manipulation of the creep lever 220. Then, the control flow proceeding to Step D19 or Step D21, the initial value α0 of the skew angle of the pump swash plate and the initial value β0 of the skew angle of the motor swash plate are set to the values which are listed for the respective forward drive range (D, L, S or M range) or the reverse drive range (R range) in FIG. 23. In either case, the vehicle speed is equal to or greater than the predetermined speed, or the throttle opening is not zero, i.e., not completely closed. Therefore, if the shift position is in any of the forward drive range, then the initial value α0 of the skew angle of the pump swash plate is set to predetermined value 5, and the initial value β0 of the skew angle of the motor swash plate is set to the maximum (β0=MAX). Likewise, if the shift position is in the reverse drive range, then the initial value α0 of the skew angle of the pump swash plate is set to predetermined value 4, and the initial value β0 of the skew angle of the motor swash plate is set to the maximum (β0=MAX).

From these target deviation α1 of the pump swash plate and the initial value α0 of the skew angle, a target skew angle α′ is calculated at Step D22 in the same way as described above. Then, a determination is made at Step D23 whether this value α′ is greater than the initial value α0, which is set at Step D19 or Step D21. If the target skew angle α′ is equal to or smaller than the initial value α0 (α′≦α0), then the target skew angle α′ is set equal to the initial value (α′=α0) at Step D24.

On the other hand, the target deviation α1 of the pump swash plate is set to zero (α1=0) at Step D6 notwithstanding the manipulation of the creep lever 220, and the initial value α0 for the pump swash plate and the initial value β0 for the motor swash plate are set in accordance with the table of FIG. 23; at Step D7. In this case, the conditions which have led the control flow to Step D6 are that the shift position is in the N or P or that the shift position is in any other range and the brake is ON. Therefore, if the shift position is in a forward drive range (D, L, S or M), then predetermined value 3 is taken as an initial value from the table. Likewise, if the shift position is in the reverse drive range (R), then predetermined value 0 is taken as an initial value. Predetermined value 2 is greater than predetermined value 3 while predetermined value 1 is greater than predetermined value 0. Thus, when the brake pedal is stepped down, a relatively small initial value is set for the skew angle of the pump swash plate. This initial value α0 and the target deviation α1 of the pump swash plate are added to achieve the target skew angle α′ for the pump swash plate at Step D8.

The following table, Table 1, describes the relations of the respective predetermined values, which are listed in FIG. 23.

TABLE 1 1) Predetermined value 5 ≧ Predetermined value 2 > Predetermined value 3 . . . (in the case of D, L, S or M) 2) Predetermined value 4 ≧ Predetermined value 1 > Predetermined value 0 . . . (in the case of R) 3) Predetermined value 3 > Predetermined value 0 4) Predetermined value 5 > Predetermined value 4

Then, a calculation is made for the deviation angle Δα (=α−α′) of this target skew angle α′ from the current skew angle α of the pump swash plate at Step D25, and a target current (target deviation current DICMDp) which is applied to the fourth linear solenoid valve 54 is set to realize this deviation angle in the pump swash plate control at Step D26. The relation between the target deviation current DICMDp and the deviation angle Δα is predetermined on the basis of the operational characteristics of the pump swash plate 32, which corresponds with the operation of the fourth linear solenoid valve 54, as shown in FIG. 25. The target deviation current DICMDp is taken from this graph in correspondence with the deviation angle Δα, which is calculated at Step D25.

With this target deviation current DICMDp, a feedback control is performed to set the skew angle of the pump swash plate to the target skew angle α′. At first, for deciding the direction for the control, a determination is made whether the current skew angle α is greater than the target skew angle α′ or not at Step D27.

If the current skew angle is greater than the target skew angle (α>α′), then the control flow proceeds to Step D29, where a determination is made whether the shift position is in the R range or not. If it is not in the R range, then a target current ICMDp is calculated at Step D31 by adding the target deviation current DICMDp to a control current IOP which is needed to retain the spool 85 of the pump control valve 84 at the neutral position. However, if the shift position is in the R range, then the target current ICMDp is calculated at Step D30 by subtracting the target deviation current DICMDp from the control current IOP which is needed to retain the spool 85 of the pump control valve 84 at the neutral position.

On the other hand, if the current skew angle is equal to or smaller than the target skew angle (α≦α′), then the control flow proceeds to Step D28, where a determination is made whether the shift position is in the R range or not. If it is in the R range, then the target current ICMDp is calculated at Sleep D31 by adding the target deviation current DICMDp to the control current IOP which is needed to retain the spool 85 of the pump control valve 84 at the neutral position. However, if the shift position is not in the R range, then the target current ICMDp is calculated at Step D30 by subtracting the target deviation current DICMDP from the control current IOP which is needed to retain the spool 85 of the pump control valve 84 at the neutral position.

At this target current ICMDp, the fourth linear solenoid valve 54 is actuated at Step D32 to control the skew angle of the pump swash plate closer to the target skew angle α′. As for the skew angle β of the motor swash plate, the initial value β0 has been set as mentioned previously. Because this initial value is either zero or the maximum value, a maximum current or zero current is set as a target current ICMDm at Step D33. At this target current, the fifth linear solenoid valve 55 is actuated at Step D34 to control the motor swash plate to the upright position or to the maximally tilted position.

It is understood from the above description that when the vehicle is in the N or P range, the skew angle of the pump swash plate is set to zero, but when the vehicle is in any other drive range, the skew angle of the pump swash plate is controlled to the target skew angle α′, and the vehicle comes into creeping condition. As the target skew angle α′ of the pump swash plate changes with the amount of manipulation of the creep lever 220, the inching of the vehicle is possible with the operation of the creep lever.

Now, the Ne control B5 is described with reference to FIG. 26. This is to control the skew angles of the swash plates of the pump and the motor and the operation of the lock-up brake, thereby regulating the engine rotational speed Ne to the target engine rotational speed Ne′ in feedback control. At first, a deviation rotational speed Δ Ne is calculated by subtracting the actual engine rotational speed Ne from the target engine rotational speed Ne′ at Step E1 (Δ Ne=Ne′−Ne).

Then, at Step E2, a determination is made which quadrant the condition of the transmission is in, of the graph of FIG. 2. As mentioned previously, this graph describes the relation of the changes in the total speed ratio to the changes in the skew angles α and β of the swash plates of the pump and the motor. For controlling the speed change, this graph representing the condition of the transmission in control is divided into four control quadrants or zones (Quadrant I-IV) as shown in FIG. 27, and a determination is made which control quadrant the actual condition is in at Step E2. This determination is described in the following with reference to FIG. 28.

As shown in FIG. 28, first a determination is made at Step F1 whether the skew angle α of the pump swash plate is greater than an intermediate angle αm, which is between zero and the maximum value. This is to determine whether the skew angle α of the pump swash plate is near the maximum skew angle in the forward drive direction α F(MAX). If the skew angle a of the pump swash plate is equal to or smaller than the intermediate angle αm (α≦αm), then the transmission is in the condition which corresponds to Quadrant I in FIG. 27, so the control flow proceeds to Step F5, whose control is described in detail in FIG. 29 (a). There, the control of the lock-up brake is turned OFF at Step F51; the control of the pump swash plate is turned ON at Step F52; and the control of the motor swash plate is turned OFF at Step F53. In this way, only the control of the pump swash plate is carried out for the condition in Quadrant I. Also, in this zone, the angle control B4 and the Ne control are switched, i.e., the controls for the start and stop of the vehicle are smoothly exchanged.

If the skew angle α of the pump swash plate is greater than the intermediate angle αm (α>αm), then a determination is made at Step F2 whether the skew angle β of the motor swash plate is greater than an intermediate angle βm, which is between zero and the maximum value. This is to determine whether the skew angle of the motor swash plate is near the maximum skew angle or not. If the skew angle of the motor swash plate is greater than an intermediate angle (β>βm), then the transmission is in the condition Which corresponds to Quadrant II in FIG. 27, so the control flow proceeds to Step F6, whose control is described in detail in FIG. 29 (b). There, the control of the lock-up brake is turned OFF at Step F61; the control of the pump swash plate is turned ON at Step F62; and the control of the motor swash plate is turned ON at Step F63. In this way, the controls of the pump and motor swash plates are carried out at the same time for the condition in Quadrant II, and these two controls are carried over from one control to the other smoothly.

If the skew angle of the motor swash plate is equal to and smaller than an intermediate angle (β≦βm), then a determination is made whether the skew angle β of the motor swash plate is zero or not at Step F3. If the skew angle β of the motor swash plate is not zero, then the transmission is in the condition which corresponds to Quadrant III in FIG. 27, so the control flow proceeds to Step F7, whose control is described in detail in FIG. 29 (c). There, the control of the lock-up brake is turned OFF at Step F71; the control of the pump swash plate is turned OFF at Step F72; and the control of the motor swash plate is turned ON at Step F73. In this way, only the control of the motor swash plate is carried out for the condition in Quadrant III, and most part of the vehicle drive control is performed in Quadrant III and Quadrant IV, which is described in the following.

If the skew angle β of the motor swash plate is zero, then a determination is made whether the vehicle speed is equal to or less than a predetermined speed or not at Step F4. If the vehicle speed is equal to or less than the predetermined speed, then the above mentioned control for Quadrant III is performed at Step F7. In this way, the lock up is not carried out for the condition where the vehicle has a speed equal to or less than the predetermined speed. However, if the vehicle speed is greater than the predetermined value, then the transmission is in the condition which corresponds to Quadrant IV in FIG. 27, so the control flow proceeds to Step F8. There, the control of the lock-up brake is turned ON at Step F81, and a determination is made of the target current ICMDL which controls the sixth linear solenoid valve 56 that actuates the lock-up brake whether this current is equal to or smaller than a predetermined value or not at Step F82. If it is equal to or smaller than the predetermined value (i.e., the engaging force of the lock-up brake is relatively small), then the control of the pump is turned OFF at Step F83; and the control of the motor is turned ON at Step F84. However, if the target current is greater than the predetermined value (i.e., the lock-up brake is engaged to a certain degree), then not only the control of the pump is turned OFF at Step F72, but also the control of the motor is turned OFF at Step F73.

After the determination of the control zone is performed at Step E2 of FIG. 26 in this way, a gain coefficient K for speed change speed is set at Step E3. As the gain coefficient K is predetermined in relation to the vehicle speed V as shown in FIG. 30, the gain coefficient K is taken in correspondence with the actual vehicle speed V at the moment. It is clear from the figure that the smaller the vehicle speed V the greater the gain coefficient. Then, this gain coefficient K is multiplied by the deviation rotational speed Δ Ne, which is calculated at Step E1, to produce a speed change speed DI (=K×Δ Ne)at Step E4.

Thereafter, determinations are made whether the controls of the pump, the motor and the lock-up brake are ON or not at Steps E5, E7 and E9. If the control of the pump is ON, then the pump swash plate is controlled at Step E6; if the control of the motor is ON, then the motor swash plate is controlled at Step E8; and if the control of the lock-up brake is ON, then the lock up is controlled at Step E10.

First, the control of the pump swash plate executed at Step E6 is described with reference to FIG. 31. The pump swash plate is controlled to swivel in the forward drive direction αF or in the rearward drive direction αR. Although these directions are opposite to each other, the pump swash plate 32 is controlled in either direction in the same manner. Therefore, the control in the forward drive direction is described here, as an example. This is to control the current applied to the fourth linear solenoid valve 54, taking the engine rotational speed as the value to be controlled. For example, if the actual engine rotational speed Ne is less than the target engine rotational speed Ne′, then the speed ratio is controlled to become greater (i.e., shifted toward LOW), which is to make the skew angle α of the pump swash plate smaller. On the other hand, if the actual engine rotational speed Ne is greater than the target engine rotational speed Ne′, then the speed ratio is controlled to become smaller (i.e., shifted toward TOP), which is to make the skew angle α of the pump swash plate greater.

In this control flow, a target deviation current DICMDp is set at Step G1 as shown in FIG. 31. The target deviation current DICMDp is predetermined in relation with the speed change speed DI as shown in FIG. 32, and from this relation, a value which corresponds with the speed change speed DI(1) that is calculated at Step E4 in FIG. 26 is taken as the target deviation current DICMDp(1). Then, a determination is made whether the target engine rotational speed Ne′ is greater than the actual engine rotational speed (i.e., the current engine rotational speed) Ne at Step G2.

If the target engine rotational speed is greater than the actual engine rotational speed (Ne′>Ne), then the control proceeds to Step G3, where a calculation is made for the target current ICMDp which is required for making the speed change ratio greater (i.e., to shift toward LOW). This target current is calculated by adding the target deviation current DICMDp to the control current IOp which is required for retaining the spool 85 of the pump control valve 84 at the neutral position. On the other hand, if the target engine rotational speed is equal to or less than the actual engine rotational speed (Ne′≦Ne), then the control proceeds to Step G4, where a calculation is made for the target current ICMDp which is required for making the speed change ratio smaller (i.e., to shift toward TOP). This target current is calculated by subtracting the target deviation current DICMDp from the control current IOp which is required for retaining the spool 85 at the neutral position.

With this target current ICMDp as a control current, the fourth linear solenoid valve 54 is actuated at Step G5. FIG. 33 shows the relation between the control pressure PC and the target current ICMDp which is applied to the linear solenoid valve 54. If the target current ICMDp equals the above mentioned control current IOp, then the pump control valve 84 is kept at the neutral. When the target current ICMDp is increased or decreased from this value, the speed is shifted toward LOW or toward TOP, respectively. In this way, the feedback control is performed to bring the actual engine rotational speed Ne close to the target engine rotational speed Ne′.

Now, the control of the motor swash plate which is carried out at Step E8 is described with reference to FIG. 34. This is to control the current applied to the fifth linear solenoid valve 55, taking the engine rotational speed as the value to be controlled. For example, if the actual engine rotational speed Ne is less than the target engine rotational speed Ne′, then the speed ratio is controlled to become greater (i.e., shifted toward LOW), which is to make the skew angle β of the motor swash plate greater. On the other hand, if the actual engine rotational speed Ne is greater than the target engine rotational speed Ne′, then the speed ratio is controlled to become smaller (i.e., shifted toward TOP), which is to make the skew angle β of the motor swash plate smaller.

In this control flow, a target deviation current DICMDm is set at Step H1 as shown in FIG. 34. The target deviation current DICMDm is predetermined in relation with the speed change speed DI as shown in FIG. 35, and from this relation, a value which corresponds with the speed change speed DI (1) that is calculated at Step E4 in FIG. 26 is taken as the target deviation current DICMDm(1). Then, a determination is made whether the target engine rotational speed Ne′ is greater than the actual engine rotational speed (i.e., the current engine rotational speed) Ne at Step H2.

If the target engine rotational speed is greater than the actual engine rotational speed (Ne′>Ne), then the control proceeds to Step H3, where a calculation is made for the target current ICMDm which is required for making the speed change ratio greater (i.e., to shift toward LOW). This target current is calculated by adding the target deviation current DICMDm to the control current IOm which is required for retaining the spool 88 of the motor control valve 87 at the neutral position. On the other hand, if the target engine rotational speed is equal to or less than the actual engine rotational speed (Ne′≦Ne), then the control proceeds to Step H4, where a calculation is made for the target current ICMDm which is required for making the speed change ratio smaller (i.e., to shift toward TOP). This target current is calculated by subtracting the target deviation current DICMDm from the control current IOm which is required for retaining the spool 88 at the neutral position.

With this target current ICMDm as a control current, the fifth linear solenoid valve 55 is actuated at Step H5. FIG. 36 shows the relation between the control pressure MC and the target current ICMDm. If the target current ICMDm equals the above mentioned control current IOm, then the motor control valve 87 is kept neutral. As the target current ICMDm is increased or decreased from this value, the speed is shifted toward LOW or toward TOP, respectively. In this way, the feedback control is performed to bring the actual engine rotational speed Ne close to the target engine rotational speed Ne′.

Now, the control of the lock-up brake which is carried out at Step E10 is described with reference to FIG. 37. This is to control the current applied to the sixth linear solenoid valve 56 by determining the strength of the lock up from the ratio of the actual engine rotational speed Ne to the target engine rotational speed Ne′. For example, if the actual engine rotational speed Ne is less than the target engine rotational speed Ne′, then the control current applied to the sixth linear solenoid valve 56 is decreased to make the lock up force smaller. On the other hand, if the actual engine rotational speed Ne is greater than the target engine rotational speed Ne′, then the control current applied to the sixth linear solenoid valve 56 is increased to make the lock up force greater.

In this control flow, a target deviation current DICMDL is set at Step I1 as shown in FIG. 37. The target deviation current DICMDL is predetermined in relation with the speed change speed DI as shown in FIG. 38, and from this relation, a value which corresponds with the speed change speed DI (1) that is calculated at Step E4 in FIG. 26 is taken as the target deviation current DICMDL(1). Then, a determination is made whether the target engine rotational speed Ne′ is greater than the actual engine rotational speed (i.e., the current engine rotational speed) Ne at Step 12.

If the target engine rotational speed is greater than the actual engine rotational speed (Ne′>Ne), then the control proceeds to Step I4, where a calculation is made for a new control current, i.e., a target current ICMDL which will be applied to the sixth linear solenoid valve 56 by subtracting the target deviation current DICMDL from the control current ICMDL that is applied presently to the six linear solenoid valve 56. Then, this target current is set to weaken the present lock-up force. On the other hand, if the target engine rotational speed is equal to or less than the actual engine rotational speed (Ne′≦Ne), then the control proceeds to Step 13, where a calculation is made for a new control current, i.e., a target current ICMDL which will be applied to the sixth linear solenoid valve 56, by adding the target deviation current DICMDL to the control current ICMDL that is applied presently to the six linear solenoid valve 56. Then, this target current is set to strengthen the present lock-up force.

With this target current ICMDL as a control current, the sixth linear solenoid valve 56 is actuated at Step I5. FIG. 39 shows the relation between the control pressure LC and the target current ICMDL, in which the control pressure LC is proportional to the target current ICMDL.

Now, the relief control at Step B6 of FIG. 15 is described. This control is to adjust the relief pressure that is applied to the high pressure relief valves 75F and 75R, which are provided on the first and second oil passages 26 a and 26 b comprising the hydraulic closed circuit 26. The pressures of the first and second oil passages 26 a and 26 b change in correspondence with the change of the vehicle drive direction (forward drive and rearward drive) and the condition of the vehicle speed (in acceleration or deceleration) as shown in FIG. 42. Therefore, this relief control responds to the change of these conditions to perform high pressure relief control. The pressure of the oil passage which is at the lower pressure is supplied from the low pressure relief valve 74 through the shuttle valve 70.

In this system, the pressure relief of the oil passage which is at the higher pressure is performed by the high pressure relief valves, which release quickly any overpressure and lower the pressure to a predetermined value, to alleviate the shock of any surge pressure. However, the target relief pressure is controlled in correspondence with the skew angle of the pump swash plate because the maximum pressure generated while the vehicle is in acceleration varies with the skew angle of the pump swash plate. Also, the target relief pressure is controlled in correspondence with the vehicle speed so that an appropriate engine brake force is realized while the vehicle is in deceleration.

The flow of this control is shown in FIGS. 40 and 41, and the circled letters, “B”, in the figures combine the two into one continuous flow. In this control, at first, a determination is made of the shift position whether it is in the N or P range at Step J1. If it is in the N or P range, then the control flow proceeds to Steps J8 and J9. However, if the shift position is in any other range than the N or P range, then a determination is made whether the engine rotational speed Ne is equal to or less than a predetermined rotational speed at Step J2. If it is equal to or less than the predetermined rotational speed, then the control flow also proceeds to Steps J8 and J9. However, if the engine rotational speed Ne is above the predetermined rotational speed, then a determination is made whether the vehicle speed V is equal to or less than a predetermined vehicle speed at Step J3. If it is above the predetermined vehicle speed, then the control flow proceeds to Steps J6 and J7. However, if the vehicle speed is equal to or less than the predetermined vehicle speed, then a determination is made of the engine throttle opening θ TH whether it is equal to or smaller than a predetermined throttle opening at Step J4. If it is above the predetermined throttle opening, then the control flow also proceeds to Steps J6 and J7. However, if the engine throttle opening is smaller than the predetermined throttle opening, then a determination is made whether the brake is ON or not at Step J5. If it is ON, then the control flow proceeds to Steps J8 and J9). However, if the brake is OFF, then the control flow proceeds to Steps J6 and J7.

A target relief pressure Pha is set in correspondence with the skew angle α of the pump swash plate as shown in FIG. 43 at Step J6. In the figure, line a1 represents the maximum allowable pressure of the apparatus, and line a2 represents the target pressure which is used to alleviate surge pressures. Furthermore, the target relief pressure Phb is set in correspondence with the vehicle speed V as shown in FIG. 44 at Step J7. This relief pressure is a target relief pressure which is set on the basis of a required engine brake force.

On the other hand, a target relief pressure can be set on the basis of the specific conditions which are determined at Steps J1-J5 as shown in FIG. 45 at Steps J8 and J9.

Specifically:

(1) if the shift position is in the N or P range (which is determined at Step J1), then the target relief pressures Pha and Phb are set to zero so that the system is made to bypass the hydraulic closed circuit to establish a neutral condition.

(2) if the engine rotational speed decreases below the idling rotational speed (i.e., a predetermined speed) (determined at Step J2), then the target relief pressure is set lower by the amount which corresponds to the rotational speed decrease to prevent the engine from stopping.

(3) if the shift position is in any drive range other than the N or P range, the engine rotational speed is above the idling rotational speed, the vehicle speed is equal to or less than the predetermined speed (a relatively low speed), the throttle opening is equal to or smaller than the predetermined throttle opening (a relatively small opening), and the brake is operated, then the condition of the vehicle is judged as being equivalent to stopping, and the target relief pressure is set to a predetermined value.

For this predetermined value for the target relief pressure, the pressures of the oil passages on both the acceleration and deceleration sides are made almost equal to each other (Pha=Phb), and this pressure is delivered by the regulator valve 60. If the condition “Pha=Phb” is not satisfied, then the shuttle valve 70 is shifted so that the pressure in the lower side is adjusted by the relief valve 74. Also, the target relief pressure may be set to a small value which will give a smooth change when the driver releases the brake pedal and steps down the accelerator pedal.

As mentioned previously, the target relief pressures Pha and Phb can be realized as desired by controlling the currents applied to the second and third linear solenoid valves 52 and 53. After the target relief pressures Pha and Phb are determined as mentioned above, the target control currents ICMDa and ICMDb which are necessary to set these relief pressures for acceleration and for deceleration respectively are set at Steps J10 and J11. The relation between the target control current and the target relief pressure is shown in FIG. 46, and from this relation, the target control current ICMD which gives the target relief pressure Pha or Phb is taken.

Then, a determination is made whether the shift position is in the R (rearward) range or not at Step J12. If the shift position is not in the R range, then the second linear solenoid valve 52, which controls the relief pressure of the first oil passage 26 a, is actuated with the target control current ICMDa for acceleration at Step J13, and the third linear solenoid valve 53, which controls the relief pressure of the second oil passage 26 b, is actuated with the target control current ICMDb for deceleration at Step J14. However, if the shift position is in the R range, then the second linear solenoid valve 52 is actuated with the target control current ICMDb for deceleration at Step J15, and the third linear solenoid valve 53 is actuated with the target control current ICMDa at Step J16.

Now, the variable notch control B7 is described with reference to FIG. 47. This is to optimize the pressure change of each cylinder bore in correspondence with the discharges of the pump 24 and the motor 25 and the high pressure generated in the operation thereof. For example, an open and close control which is shown in the table of FIG. 48 is performed for the operation of the pump 24.

In this control flow, at first, the displacement volume Vp of the pump is determined from the skew angle α of the pump swash plate at Step K1. FIG. 49 shows the relation between the skew angle of the pump swash plate and the displacement volume of the pump, and it is calculated by Equation (2). Then, the discharge of the pump Qp (=Vp×Np) is calculated at Step K2. Here, Np is the rotational speed of the pump, and it is calculated by Equation (3) with the engine rotational speed Ne and the rotational speed of the wheels Nv, which values are detected with the respective sensors.

Vp=S×PD× tan α×N  (2)

Here, S is the sectional area of the cylinder bore; PD is the diameter of the pitch circle of the cylinders; and N is the number of cylinders.

NP={Ne−i1×i2×(1+ip)×Nv}/ip  (3)

Here, i1 is the reduction ratio of the power transmission mechanism 4; i2 is the reduction ratio of the final speed reduction mechanism 5; and ip is the reduction ratio of the power distribution mechanism 3 (refer to FIG. 50).

Then, a determination is made whether the discharge of the pump Qp is greater than a predetermined value Qp0 or not at Step K3. If the discharge is greater than the predetermined value (Qp>Qp0), then the control flow proceeds to Steps K7 and K8. However, if the discharge is equal to or less than the predetermined value (Qp≦Qp0), then another determination is made of the high pressure Ph generated in the oil passage which is at the higher pressure whether this high pressure is greater than a predetermined pressure Ph0 or not at Step K4. If it is so (Ph>Ph0), then the control flow proceeds to Steps K5 and K6. However, if the high pressure Ph is equal to or smaller than the predetermined pressure Ph0 (Ph≦Ph0), then the control flow proceeds to Steps K7 and K8. The control at Step k5 is to close the variable notch valves, so the solenoid valves 45 and 46 are turned ON at Step K6. On the other hand, the control at Step k7 is to open the variable notch valves, so the solenoid valves 45 and 46 are turned OFF at Step K8.

Now, a description is made of the adjustment of the hydraulic pressures which are used for the actuation of the servo cylinders 92 a and 92 b that swivel the pump swash plate 32 and the servo cylinders 96 a and 96 b that swivel the motor swash plate 38. The hydraulic pressures required for the actuation of the servo cylinders 92 and 96 are not always constant. They need to be adjusted in proportion to the respective moment forces (these forces refer to the periodic forces which act around the trunnions of the respective swash plates 32 and 38 while the hydraulic pump 24 and the hydraulic motor 25 rotate with the pistons thereof reciprocating inside the cylinders, and the inside pressures thereof vary accordingly). Therefore, the hydraulic pressures are adjusted in correspondence with the magnitudes of the moment forces, which act on the swash plates 32 and 38, and the moment forces are estimated in accordance with a few conditions. As the varying moment forces affect mostly the pressure Ph of the oil passage which is at the higher pressure, the moment forces are estimated from the pressure Ph of the oil passage which is at the higher pressure.

On the other hand, the hydraulic pressures of the servo cylinders 92 and 96 are supplied from the line pressure PL which is adjusted by the regulator valves 60 as mentioned previously. Furthermore, the regulator valve 60 is controlled in correspondence with the amount of actuation of the first linear solenoid valve 51, which is driven electromagnetically. Therefore, by controlling the amount of actuation of the first linear solenoid valve 51, the hydraulic pressures of the servo cylinders 92 and 96 can be controlled.

Thus, the estimation of the moment forces which act on the swash plates 32 and 38 and the adjustment of the hydraulic pressures of the servo cylinders 92 and 96 on the basis of these moment forces are to detect the pressure Ph of the oil passage which is at the higher pressure and to set the amount of actuation of the first linear solenoid valve 51 on the basis of this estimation. Now, with reference to the flowchart of FIG. 51, a description is made of the actuation control B8 of the first linear solenoid valve 51 on the basis of the hydraulic pressure Ph. First, the pressure Ph of the oil passage which is at the higher pressure is detected by the pressure sensor 208, which detects the pressure P1 of the first oil passage 26 a of the hydraulic closed circuit 26, and the pressure sensor 209, which detects the pressure P2 of the second oil passage 26 b. This detected result is input into the control unit ECU, and the target current ICMDr is determined at Step L1 in accordance with the relation shown in FIG. 52. Then, the target current ICMDr is output from the control unit ECU for the actuation of the first solenoid valve 51 at Step L2. As shown in FIG. 52, the greater the pressure Ph of the oil passage which is at the higher pressure, the greater the target current ICMDr.

By performing this control, the line pressure PL is adjustable in correspondence with the pressure Ph of the oil passage which is at the higher pressure in a desired way as shown in FIG. 53a (FIG. 53b or FIG. 53c). The relation therebetween is predetermined such that the line pressure PL is always at the minimum pressure which is necessary for the actuation of the servo cylinders 92 and 96. Thereby, the hydraulic pressure (i.e., the line pressure PL) is not kept unnecessarily high so as to minimize the load of the prime motor, i.e., the engine, for a better fuel efficiency. Also, the adjustment of the hydraulic pressure can be performed accurately because the estimation of the moment forces is carried out in correspondence with the pressure Ph of the oil passage at the higher pressure which is most affected from the varying moment forces.

In this way, the estimation of the moment forces can be carried out from the pressure Ph of the oil passage which is at the higher pressure in the hydraulic closed circuit 26. However, it is known that even though the pressure Ph is constant, (1) if the types of the valve plates used for the swash plates 32 and 38 differ, (2) if the skew angle α of the pump swash plate and the skew angle β of the motor swash plate differ, (3) if the rotational speed Np of the hydraulic pump and the rotational speed Nm of the hydraulic motor differ, or (4) if the operational condition of the transmission differs (depending on whether it is in acceleration or in deceleration, i.e., whether the external driving force acting on the transmission comes through the hydraulic pump or through the hydraulic motor), then the moment forces vary a little. Therefore, the target current ICMDr which is calculated as described above is first set as a basic target current ICMDr0, and preferably, this basic target current is corrected in consideration of the above four factors, (1)−(4), to achieve the target current ICMDr, which activates the first linear solenoid valve 51 (specifically, this correction is performed by multiplying the basic target current ICMDr0 by correction factors). This correction procedure is described below with reference to FIGS. 54 through 59.

FIG. 54 shows the control flow B8′ through which the target current ICMDr is calculated. First, a description is made for a case in which the valve plate used is a pre-compression and pre-expansion type valve plate such as the valve plate 150 shown in FIG. 11. At first, the basic target current ICMDr0 is calculated at Step M1 from the pressure Ph of the oil passage which is detected the higher in accordance with the relation shown in FIG. 55. Then, a determination is made whether the vehicle is in acceleration or in deceleration by comparing the pressure P1 of the first oil passage 26 a with the pressure P2 of the second oil passage 26 b at Step M2. If the vehicle is in acceleration (i.e., the driving force external to the transmission acts on the hydraulic pump 24, i.e., P1≦P2), then the control flow proceeds to Step M3 and Step M4. On the other hand, if the vehicle is in deceleration (i.e., the external driving force acts on the hydraulic motor 25, i.e., P1 <P2), then the control flow proceeds to Step M3′ and Step M4′. As a result, the correction factor L1 for the skew angle α of the pump swash plate and the correction factor L2 for the skew angle β of the motor swash plate are determined, respectively.

In the case where the valve plate used is a pre-compression and pre-expansion type valve plate 150, it is known that the moment forces which act on the pump swash plate 32 and the motor swash plate 38 are: (1) greater correspondingly for smaller skew angles α of the pump swash plate and for larger skew angles β of the motor swash plate while the vehicle is in acceleration (i.e., the driving force external to the transmission acts on the hydraulic pump), and (2) greater correspondingly for larger skew angles α of the pump swash plate and for smaller skew angles β of the motor swash plate while the vehicle is in deceleration (i.e., the driving force external to the transmission acts on the hydraulic motor). Therefore, if the vehicle is in acceleration, then the correction factor L1 for the skew angle α of the pump swash plate is determined in accordance with the relation shown in FIG. 56a, and the correction factor L2 for the skew angle β of the motor swash plate is determined in accordance with the relation shown in FIG. 57a, respectively. On the other hand, if the vehicle is in deceleration, then the correction factor L1 is determined in accordance with the relation shown in FIG. 56b, and the correction factor L2 is determined in accordance with the relation shown in FIG. 57b.

After the correction factors L1 and L2 are determined, the control flow proceeds to Step M5 and Step M6 to determine the correction factor L3 which concerns the rotational speed Np of the hydraulic pump and the correction factor L4 which concerns the rotational speed Nm of the hydraulic motor. It is known that when the pre-compression and pre-expansion type valve plate 150 is used, the moment forces which act on the pump swash plate 32 and the motor swash plate 38 are greater correspondingly for higher rotational speeds Np of the hydraulic pump and for higher rotational speeds Nm of the hydraulic motor notwithstanding whether the vehicle is in acceleration or deceleration. Therefore, the correction factor L3 is determined in accordance with the relation shown in FIG. 58a, and the correction factor L4 is determined in accordance with the relation shown in FIG. 59a, respectively.

The calculation of the target current ICMDr is performed by the control unit ECU. At Step M7, the calculated basic target current ICMDr0 is multiplied by these correction factors L1-L4 to achieve a target current ICMDr, and this target current ICMDr is provided from the control unit to the first linear solenoid valve 51 for the actuation thereof at Step M8. FIG. 60 shows the relation between the target current ICMDr and the line pressure PL, which is controlled by this current. In this way, the target current ICMDr is achieved by correcting the basic target current ICMDr0, which is determined from the pressure Ph of the oil passage at the higher pressure in the hydraulic circuit. Therefore, the precision of the hydraulic control of the servo cylinders 92 and 96 is improved.

Now, a description is made of a case where the valve plate is a lead angle type valve plate 160 such as shown in FIG. 61. This valve plate 160 is provided with a first port 161, which is connected to the second oil passage 26 b, and a second port 162, which is connected to the first oil passage 26 a. Furthermore, an additional oil passage 161 a is provided in fluid communication with this first port 161, and another additional oil passage 162 a is provided in fluid communication with this second port 162. These additional oil passages 161 a and 162 a are provided extending across the top dead center (T.D.C.) and the bottom dead center (B.D.C.), respectively, each having an orifice which opens at a location between the first and second ports 161 and 162.

In the case where this lead angle type valve plate 160 is used, it is known that the moment forces which act on the pump swash plate 32 and the motor swash plate 38 are: (1) greater correspondingly for larger skew angles α of the pump swash plate and for smaller skew angles β of the motor swash plate while the vehicle is in acceleration (i.e., the driving force external to the transmission acts on the hydraulic pump), and (2) greater correspondingly for smaller skew angles a of the pump swash plate and for larger skew angles β of the motor swash plate while the vehicle is in deceleration (i.e., the external driving force acts on the hydraulic motor). Therefore, in the control flow shown in FIG. 54, if the vehicle is in acceleration (i.e., P1≦P2), then the correction factor L1 for the skew angle α of the pump swash plate is determined in accordance with the relation shown in FIG. 56b, and the correction factor L2 for the skew angle b of the motor swash plate is determined in accordance with the relation shown in FIG. 57b, respectively. On the other hand, if the vehicle is in deceleration (P1<P2), then the correction factor L1 is determined in accordance with the relation shown in FIG. 56a, and the correction factor L2 is determined in accordance with the relation shown in FIG. 57a.

It is also known that when the lead angle type valve plate 160 is used, the moment forces which act on the pump swash plate 32 and the motor swash plate 38 are greater correspondingly for lower rotational speeds Np of the hydraulic pump and for lower rotational speeds Nm of the hydraulic motor notwithstanding whether the vehicle is in acceleration or in deceleration. Therefore, in the control flow shown in FIG. 54, the correction factor L3 is determined in accordance with the relation shown in FIG. 58b, and the correction factor L4 is determined in accordance with the relation shown in FIG. 59b, respectively.

In the above embodiment, the procedure for calculating the target current ICMDr is performed first by determining a basic target current ICMDr0 from the moment forces estimated from the pressure Ph of the oil passage which is at the higher pressure and then by correcting this basic target current ICMDr0 in consideration of the above four factors (1)-(4). However, another procedure may be also applied in which the estimated moment forces are corrected first in accordance with the level of the moments in consideration of the above four factors (1)-(4), and then, the target current ICMDr is determined from these corrected moment forces.

The invention being thus described, it will be obvious that the same may be varied in many ways. Such variations are not to be regarded as a departure from the spirit and scope of the invention, and all such modifications as would be obvious to one skilled in the art are intended to be included within the scope of the following claims.

RELATED APPLICATIONS

The application claims the priority of Japanese Patent Application No. 10-008555 filed on Jan. 20, 1998, which is incorporated herein by reference. 

What is claimed is:
 1. A hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, said hydraulic pump driven by a prime motor delivering oil that actuates said hydraulic motor into rotation; said hydraulic controller comprising: hydraulically actuated swash plate tilting means which tilts a swash plate of at least one of said hydraulic pump and said hydraulic motor around an axis for capacity control; pressure controlling means which adjusts a hydraulic pressure used for actuating said swash plate tilting means; moment estimating means which estimates a moment force that acts on said swash plate from said at least one of said hydraulic pump and said hydraulic motor; and a hydraulic closed circuit comprising a first oil passage and a second oil passage; wherein: said pressure controlling means adjusts said hydraulic pressure correspondingly higher when said moment force estimated by said moment estimating means becomes greater, and said moment estimating means comprises first pressure detecting means which detects the pressure of said first oil passage and second pressure detecting means which detects the pressure of said second oil passage; and said moment estimating means estimates said moment force from the pressure which is higher of said two pressures of said first oil passage and said second oil passage detected by said first pressure detecting means and said second pressure detecting means, wherein valve plates incorporated in said hydraulic pump and said hydraulic motor are of a pre-compression and pre-expansion type, and wherein  said moment estimating means estimates said moment force as being correspondingly greater for smaller skew angles of the swash plate of said hydraulic pump and for larger skew angles of the swash plate of said hydraulic motor while the transmission is in acceleration, in which said hydraulic motor is driven by said hydraulic pump.
 2. A hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, said hydraulic pump driven by a prime motor delivering oil that actuates said hydraulic motor into rotation; said hydraulic controller comprising: hydraulically actuated swash plate tilting means which tilts a swash plate of at least one of said hydraulic pump and said hydraulic motor around an axis for capacity control; pressure controlling means which adjusts a hydraulic pressure used for actuating said swash plate tilting means; moment estimating means which estimates a moment force that acts on said swash plate from said at least one of said hydraulic pump and said hydraulic motor; and a hydraulic closed circuit comprising a first oil passage and a second oil passage; wherein: said pressure controlling means adjusts said hydraulic pressure correspondingly higher when said moment force estimated by said moment estimating means becomes greater, and said moment estimating means comprises first pressure detecting means which detects the pressure of said first oil passage and second pressure detecting means which detects the pressure of said second oil passage; and said moment estimating means estimates said moment force from the pressure which is higher of said two pressures of said first oil passage and said second oil passage detected by said first pressure detecting means and said second pressure detecting means, wherein valve plates incorporated in said hydraulic pump and said hydraulic motor are of a pre-compression and pre-expansion type, and wherein  said moment estimating means estimates said moment force as being correspondingly greater for larger skew angles of the swash plate of said hydraulic pump and for smaller skew angles of the swash plate of said hydraulic motor while the transmission is in deceleration, in which said hydraulic pump is driven by said hydraulic motor.
 3. A hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, said hydraulic pump driven by a prime motor delivering oil that actuates said hydraulic motor into rotation; said hydraulic controller comprising: hydraulically actuated swash plate tilting means which tilts a swash plate of at least one of said hydraulic pump and said hydraulic motor around an axis for capacity control; pressure controlling means which adjusts a hydraulic pressure used for actuating said swash plate tilting means; moment estimating means which estimates a moment force that acts on said swash plate from said at least one of said hydraulic pump and said hydraulic motor; and a hydraulic closed circuit comprising a first oil passage and a second oil passage; wherein: said pressure controlling means adjusts said hydraulic pressure correspondingly higher when said moment force estimated by said moment estimating means becomes greater, and said moment estimating means comprises first pressure detecting means which detects the pressure of said first oil passage and second pressure detecting means which detects the pressure of said second oil passage; and said moment estimating means estimates said moment force from the pressure which is higher of said two pressures of said first oil passage and said second oil passage detected by said first pressure detecting means and said second pressure detecting means, wherein valve plates used in said hydraulic pump and said hydraulic motor are of a lead angle type, and wherein  said moment estimating means estimates said moment force as being correspondingly greater for larger skew angles of the swash plate of said hydraulic pump and for smaller skew angles of the swash plate of said hydraulic motor while the transmission is in acceleration, in which said hydraulic motor is driven by said hydraulic pump.
 4. A hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, said hydraulic pump driven by a prime motor delivering oil that actuates said hydraulic motor into rotation; said hydraulic controller comprising: hydraulically actuated swash plate tilting means which tilts a swash plate of at least one of said hydraulic pump and said hydraulic motor around an axis for capacity control; pressure controlling means which adjusts a hydraulic pressure used for actuating said swash plate tilting means; moment estimating means which estimates a moment force that acts on said swash plate from said at least one of said hydraulic pump and said hydraulic motor; and a hydraulic closed circuit comprising a first oil passage and a second oil passage; wherein: said pressure controlling means adjusts said hydraulic pressure correspondingly higher when said moment force estimated by said moment estimating means becomes greater, and said moment estimating means comprises first pressure detecting means which detects the pressure of said first oil passage and second pressure detecting means which detects the pressure of said second oil passage; and said moment estimating means estimates said moment force from the pressure which is higher of said two pressures of said first oil passage and said second oil passage detected by said first pressure detecting means and said second pressure detecting means, wherein valve plates used in said hydraulic pump and said hydraulic motor are of a lead angle type, and wherein  said moment estimating means estimates said moment force as being correspondingly greater for smaller skew angles of the swash plate of said hydraulic pump and for larger skew angles of the swash plate of said hydraulic motor while the transmission is in deceleration, in which said hydraulic pump is driven by said hydraulic motor.
 5. A hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, said hydraulic pump driven by a prime motor delivering oil that actuates said hydraulic motor into rotation; said hydraulic controller comprising: hydraulically actuated swash plate tilting means which tilts a swash plate of at least one of said hydraulic pump and said hydraulic motor around an axis for capacity control; pressure controlling means which adjust a hydraulic pressure used for actuating said swash plate tilting means; and a hydraulic closed circuit comprising a first oil passage and a second oil passage; wherein: said pressure controlling means adjusts said hydraulic pressure correspondingly higher when a moment force that acts on said swash plate from said at least one of said hydraulic pump and said hydraulic motor becomes greater, and said pressure controlling means adjusts said hydraulic pressure correspondingly higher as the pressure which is higher of said two pressures of said first oil passage and said second oil passage detected by said first pressure detecting means and said second pressure detecting means becomes higher, wherein valve plates incorporated in said hydraulic pump and said hydraulic motor are of a pre-compression and pre-expansion type, and wherein  said pressure controlling means adjusts said hydraulic pressure correspondingly higher for smaller skew angles of the swash plate of said hydraulic pump an for larger skew angles of the swash plate of said hydraulic motor while the transmission is in acceleration, in which said hydraulic motor is driven by said hydraulic pump.
 6. A hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, said hydraulic pump driven by a prime motor delivering oil that actuates said hydraulic motor into rotation; said hydraulic controller comprising: hydraulically actuated swash plate tilting means which tilts a swash plate of at least one of said hydraulic pump and said hydraulic motor around an axis for capacity control; pressure controlling means which adjust a hydraulic pressure used for actuating said swash plate tilting means; and a hydraulic closed circuit comprising a first oil passage and a second oil passage; wherein: said pressure controlling means adjusts said hydraulic pressure correspondingly higher when a moment force that acts on said swash plate from said at least one of said hydraulic pump and said hydraulic motor becomes greater, and said pressure controlling means adjusts said hydraulic pressure correspondingly higher as the pressure which is higher of said two pressures of said first oil passage and said second oil passage detected by said first pressure detecting means and said second pressure detecting means becomes higher, wherein valve plates incorporated in said hydraulic pump and said hydraulic motor are of a pre-compression and pre-expansion type, and wherein  said pressure controlling means adjusts said hydraulic pressure correspondingly higher for larger skew angles of the swash plate of said hydraulic pump an for smaller skew angles of the swash plate of said hydraulic motor while the transmission is in deceleration, in which said hydraulic pump is driven by said hydraulic motor.
 7. A hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, said hydraulic pump driven by a prime motor delivering oil that actuates said hydraulic motor into rotation; said hydraulic controller comprising: hydraulically actuated swash plate tilting means which tilts a swash plate of at least one of said hydraulic pump and said hydraulic motor around an axis for capacity control; pressure controlling means which adjust a hydraulic pressure used for actuating said swash plate tilting means; and a hydraulic closed circuit comprising a first oil passage and a second oil passage; wherein: said pressure controlling means adjusts said hydraulic pressure correspondingly higher when a moment force that acts on said swash plate from said at least one of said hydraulic pump and said hydraulic motor becomes greater, and said pressure controlling means adjusts said hydraulic pressure correspondingly higher as the pressure which is higher of said two pressures of said first oil passage and said second oil passage detected by said first pressure detecting means and said second pressure detecting means becomes higher, wherein valve plates used in said hydraulic pump and said hydraulic motor are of a lead angle type, and wherein  said pressure controlling means adjusts said hydraulic pressure correspondingly higher for larger skew angles of the swash plate of said hydraulic pump and for smaller skew angles of the swash plate of said hydraulic motor while the transmission is in acceleration, in which said hydraulic motor is driven by said hydraulic pump.
 8. A hydraulic controller for a variable capacity hydraulic transmission which comprises a hydraulic pump and a hydraulic motor, said hydraulic pump driven by a prime motor delivering oil that actuates said hydraulic motor into rotation; said hydraulic controller comprising: hydraulically actuated swash plate tilting means which tilts a swash plate of at least one of said hydraulic pump and said hydraulic motor around an axis for capacity control; pressure controlling means which adjust a hydraulic pressure used for actuating said swash plate tilting means; and a hydraulic closed circuit comprising a first oil passage and a second oil passage; wherein: said pressure controlling means adjusts said hydraulic pressure correspondingly higher when a moment force that acts on said swash plate from said at least one of said hydraulic pump and said hydraulic motor becomes greater, and said pressure controlling means adjusts said hydraulic pressure correspondingly higher as the pressure which is higher of said two pressures of said first oil passage and said second oil passage detected by said first pressure detecting means and said second pressure detecting means becomes higher, wherein valve plates used in said hydraulic pump and said hydraulic motor are of a lead angle type, and wherein  said pressure controlling means adjusts said hydraulic pressure correspondingly higher for smaller skew angles of the swash plate of said hydraulic pump and for larger skew angles of the swash plate of said hydraulic motor while the transmission is in deceleration, in which said hydraulic pump is driven by said hydraulic motor. 